Transferring heat between fluids

ABSTRACT

Heat exchangers can include heat exchange plates with: front and back exterior surfaces exposed to a non-working fluid; an interior working fluid flow channel between the front and back exterior surfaces, comprising a first plurality of parallel flow paths in a first direction and a second plurality of parallel flow paths in a second direction.

TECHNICAL FIELD

This invention relates to transferring heat between fluids and, morespecifically, to transferring heat between fluids using heat exchangeplates.

BACKGROUND

Energy consumption and demand throughout the world has grown at anexponential rate. This demand is expected to continue to rise,particularly in developing countries in Asia and Latin America. At thesame time, traditional sources of energy, namely fossil fuels, are beingdepleted at an accelerating rate and the cost of exploiting fossil fuelscontinues to rise. Environmental and regulatory concerns areexacerbating that problem.

Solar-related renewable energy is one alternative energy source that mayprovide a portion of the solution to the growing demand for energy.Solar-related renewable energy is appealing because, unlike fossilfuels, uranium, or even thermal “green” energy, there are few or noclimatic risks associated with its use. In addition, solar relatedenergy is free and vastly abundant.

Ocean Thermal Energy Conversion (“OTEC”) is a manner of producingrenewable energy using solar energy stored as heat in the oceans'tropical regions. Tropical oceans and seas around the world offer aunique renewable energy resource. In many tropical areas (betweenapproximately 20° north and 20° south latitude), the temperature of thesurface sea water remains nearly constant. To depths of approximately100 ft the average surface temperature of the sea water variesseasonally between 75° F. and 85° F. or more. In the same regions, deepocean water (between 2500 ft and 4200 ft or more) remains a fairlyconstant 40° F. Thus, the tropical ocean structure offers a large warmwater reservoir at the surface and a large cold water reservoir atdepth, with a temperature difference between the warm and coldreservoirs of between 35° F. to 45° F. This temperature difference (ΔT)remains fairly constant throughout the day and night, with smallseasonal changes.

The OTEC process uses the temperature difference between surface anddeep sea tropical waters to drive a heat engine to produce electricalenergy. OTEC power generation was identified in the late 1970's as apossible renewable energy source having a low to zero carbon footprintfor the energy produced. An OTEC power plant, however, has a lowthermodynamic efficiency compared to more traditional, high pressure,high temperature power generation plants. For example, using the averageocean surface temperatures between 80° F. and 85° F. and a constant deepwater temperature of 40° F., the maximum ideal Carnot efficiency of anOTEC power plant will be 7.5 to 8%. In practical operation, the grosspower efficiency of an OTEC power system has been estimated to be abouthalf the Carnot limit, or approximately 3.5 to 4.0%. Additionally,analysis performed by leading investigators in the 1970's and 1980's,and documented in “Renewable Energy from the Ocean, a Guide to OTEC”William Avery and Chih Wu, Oxford University Press, 1994 (incorporatedherein by reference), indicates that between one quarter to one half (ormore) of the gross electrical power generated by an OTEC plant operatingwith a ΔT of 40° F. would be required to run the water and working fluidpumps and to supply power to other auxiliary needs of the plant. On thisbasis, the low overall net efficiency of an OTEC power plant convertingthe thermal energy stored in the ocean surface waters to net electricenergy has not been a commercially viable energy production option.

An additional factor resulting in further reductions in overallthermodynamic efficiency is the loss associated with providing necessarycontrols on the turbine for precise frequency regulation. Thisintroduces pressure losses in the turbine cycle that limit the work thatcan be extracted from the warm sea water. The resulting net plantefficiency would then be between 1.5% and 2.0%

This low OTEC net efficiency compared with efficiencies typical of heatengines that operate at high temperatures and pressures has led to thewidely held assumption by energy planners that OTEC power is too costlyto compete with more traditional methods of power production.

Indeed, the parasitic electrical power requirements are particularlyimportant in an OTEC power plant because of the relatively smalltemperature difference between the hot and cold water. To achievemaximum heat transfer between the warm sea water and the working fluid,and between the cold sea water and the working fluid large heat exchangesurface areas are required, along with high fluid velocities. Increasingany one of these factors can increase the parasitic load on the OTECplant, thereby decreasing net efficiency. An efficient heat transfersystem that maximizes the energy transfer in the limited temperaturedifferential between the sea water and the working fluid would increasethe commercial viability of an OTEC power plant.

In addition to the relatively low efficiencies with seemingly inherentlarge parasitic loads, the operating environment of OTEC plants presentsdesign and operating challenges that also decrease the commercialviability of such operations. As previously mentioned, the warm waterneeded for the OTEC heat engine is found at the surface of the ocean, toa depth of 100 ft or less. The constant source of cold water for coolingthe OTEC engine is found at a depth of between 2700 ft and 4200 ft ormore. Such depths are not typically found in close proximity topopulation centers or even land masses. An offshore power plant isrequired.

Whether the plant is floating or fixed to an underwater feature, a longcold water intake pipe of 2000 ft or longer is required. Moreover,because of the large volume of water required in commercially viableOTEC operations, the cold water intake pipe requires a large diameter(typically between 6 and 35 feet or more). Suspending a large diameterpipe from an offshore structure presents stability, connection andconstruction challenges which have previously driven OTEC costs beyondcommercial viability.

Additionally, a pipe having significant length to diameter ratio that issuspended in a dynamic ocean environment can be subjected to temperaturedifferences and varying ocean currents along the length of the pipe.Stresses from bending and vortex shedding along the pipe also presentchallenges. And surface influences such as wave action present furtherchallenges with the connection between the pipe and floating platform. Acold water pipe intake system having desirable performance, connection,and construction consideration would increase the commercial viabilityof an OTEC power plant.

Environmental concerns associated with an OTEC plant have also been animpediment to OTEC operations. Traditional OTEC systems draw in largevolumes of nutrient rich cold water from the ocean depths and dischargethis water at or near the surface. Such discharge can effect, in apositive or adverse manner, the ocean environment near the OTEC plant,impacting fish stocks and reef systems that may be down current from theOTEC discharge.

SUMMARY

In some aspects, power generation plant uses ocean thermal energyconversion processes as a power source.

Further aspects relate to an offshore OTEC power plant having improvedoverall efficiencies with reduced parasitic loads, greater stability,lower construction and operating costs, and improved environmentalfootprint. Other aspects include large volume water conduits that areintegral with the floating structure. Modularity and compartmentation ofthe multi-stage OTEC heat engine reduces construction and maintenancecosts, limits off-grid operation and improves operating performance.Still further aspects provide for a floating platform havingstructurally integrated heat exchange compartments and provides for lowmovement of the platform due to wave action. The integrated floatingplatform may also provide for efficient flow of the warm water or coolwater through the multi-stage heat exchanger, increasing efficiency andreducing the parasitic power demand. Associated system can promote anenvironmentally neutral thermal footprint by discharging warm and coldwater at appropriate depth/temperature ranges. Energy extracted in theform of electricity reduces the bulk temperature to the ocean.

Further aspects relate to a floating, low heave OTEC power plant havinga high efficiency, multi-stage heat exchange system, wherein the warmand cold water supply conduits and heat exchanger cabinets arestructurally integrated into the floating platform or structure of thepower plant.

In some aspects, heat exchange plates include: front and back exteriorsurfaces exposed to a non-working fluid; and an interior working fluidflow channel between the front and back exterior surfaces, comprising afirst plurality of parallel flow paths in a first direction and a secondplurality of parallel flow paths in a second direction. Embodiments ofthese systems can include one or more of the following features.

In some embodiments, the first direction is opposite and parallel to thesecond direction.

In some embodiments, the first and second directions of the first andsecond flow paths are perpendicular to the direction of flow of thenon-working fluid.

In some embodiments, wherein the plate further comprises a first regionof relatively high working fluid mass flux when a working fluid has alow vapor quality and a second region of relatively low working fluidmass flux when the working fluid has a high vapor quality.

In some embodiments, the interior flow channel has an area of varyingspace in fluid contact with the first plurality of parallel flow pathsand the second plurality of parallel flow paths. In some cases, theplates also include one or more structural walls within the first andsecond plurality of flow paths, the one or more walls being generallyparallel to the flow path and terminating in the area of varying space.In some cases, the one or more structural walls comprises directionalvane at the terminal end and oriented in the direction of flow. In somecases, the one or more structural walls comprise a directional vane atthe proximal end of the structural wall.

In some embodiments, a flow path of the first and second plurality offlow paths comprises a void having a cross sectional area of between 155mm and 60 mm.

In some embodiments, the heat exchange plate is a composite blow moldedplate.

In some embodiments, the heat exchange plate is aluminum.

In some embodiments, the working fluid pressure drop across the plate isabout 0.2 psi/ft.

In some embodiments, the non-working fluid heat transfer coefficientranges from 900 to 1400 Btu/ft2 Rhr.

In some embodiments, the pattern of a first plurality of parallel flowpaths in a first direction and a second plurality of parallel flow pathsin a second direction repeats across the length of the heat exchangeplate. In some cases, the number of flow paths in the first and secondplurality of flow paths increases as the pattern repeats across thelength of the heat exchange plate. In some cases, the number of flowpaths in the first plurality of flow paths increases from four flowpaths per first direction to six flow paths per first direction. In somecases, the number of flow paths in the first plurality of flow pathsincreases from two flow paths per first direction to four flow paths perfirst direction.

In some embodiments, the non-working fluid is sea water.

In some embodiments, the working fluid is ammonia.

In some embodiments, the plate is an OTEC heat exchange plate.

Still further aspects include a floating ocean thermal energy conversionpower plant. A low heave structure, such as a spar, or modifiedsemi-submersible offshore structure may comprise a first deck portionhaving structurally integral warm sea water passages, multi-stage heatexchange surfaces, and working fluid passages, wherein the first deckportion provides for the evaporation of the working fluid. A second deckportion is also provided having structurally integral cold sea waterpassages, multi-stage heat exchange surfaces, and working fluidpassages, wherein the second deck portion provides a condensing systemfor condensing the working fluid from a vapor to a liquid. The first andsecond deck working fluid passages are in communication with a thirddeck portion comprising one or more vapor turbine driven electricgenerators for power generation.

In one aspect, an offshore power generation structure is providedcomprising a submerged portion. The submerged portion further comprisesa first deck portion comprising an integral multi-stage evaporatorsystem, a second deck portion comprising an integral multi-stagecondensing system; a third deck portion housing power generation andtransformation equipment; a cold water pipe and a cold water pipeconnection.

In a further aspect, the first deck portion further comprises a firststage warm water structural passage forming a high volume warm waterconduit. The first deck portion also comprises a first stage workingfluid passage arranged in cooperation with the first stage warm waterstructural passage to warm a working fluid to a vapor. The first deckportion also comprises a first stage warm water discharge directlycoupled to a second stage warm water structural passage. The secondstage warm water structural passage forms a high volume warm waterconduit and comprises a second stage warm water intake coupled to thefirst stage warm water discharge. The arrangement of the first stagewarm water discharge to the second stage warm water intake provides lowpressure loss in the warm water flow between the first and second stage.The first deck portion also comprises a second stage working fluidpassage arranged in cooperation with the second stage warm waterstructural passage to warm the working fluid to a vapor. The first deckportion also comprises a second stage warm water discharge.

In a further aspect, the submerged portion further comprises a seconddeck portion comprising a first stage cold water structural passageforming a high volume cold water conduit. The first stage cold waterpassage further comprises a first stage cold water intake. The seconddeck portion also comprises a first stage working fluid passage incommunication with the first stage working fluid passage of the firstdeck portion. The first stage working fluid passage of the second deckportion in cooperation with the first stage cold water structuralpassage cools the working fluid to a liquid. The second deck portionalso comprises a first stage cold water discharge directly coupled to asecond stage cold water structural passage forming a high volume coldwater conduit. The second stage cold water structural passage comprisesa second stage cold water intake. The first stage cold water dischargeand the second stage cold water intake are arranged to provide lowpressure loss in the cold water flow from the first stage cold waterdischarge to the second stage cold water intake. The second deck portionalso comprises a second stage working fluid passage in communicationwith the second stage working fluid passage of the first deck portion.The second stage working fluid passage in cooperation with the secondstage cold water structural passage cool the working fluid within thesecond stage working fluid passage to a liquid. The second deck portionalso comprises a second stage cold water discharge.

In a further aspect, the third deck portion may comprise a first andsecond vapor turbine, wherein the first stage working fluid passage ofthe first deck portion is in communication with the first turbine andthe second stage working fluid passage of the first deck portion is incommunication with the second turbine. The first and second turbine canbe coupled to one or more electric generators.

In still further aspects, an offshore power generation structure isprovided comprising a submerged portion, the submerged portion furthercomprises a four stage evaporator portion, a four stage condenserportion, a four stage power generation portion, a cold water pipeconnection, and a cold water pipe.

In one aspect, the four stage evaporator portion comprises a warm waterconduit including, a first stage heat exchange surface, a second stageheat exchange surface, a third stage heat exchange surface, and fourthstage heat exchange surface. The warm water conduit comprises a verticalstructural member of the submerged portion. The first, second, third andfourth heat exchange surfaces are in cooperation with first, second,third and fourth stage portions of a working fluid conduit, wherein aworking fluid flowing through the working fluid conduit is heated to avapor at each of the first, second, third, and fourth stage portions.

In one aspect, the four stage condenser portion comprises a cold waterconduit including a first stage heat exchange surface, a second stageheat exchange surface, a third stage heat exchange surface, and fourthstage heat exchange surface. The cold water conduit comprises a verticalstructural member of the submerged portion. The first, second, third andfourth heat exchange surfaces are in cooperation with first, second,third and fourth stage portions of a working fluid conduit, wherein aworking fluid flowing through the working fluid conduit is cooled to aliquid at each of the first, second, third, and fourth stage portions,with a progressively higher temperature at each successive stage.

In yet another aspect, first, second, third and fourth stage workingfluid conduits of the evaporator portion are in communication withfirst, second, third and fourth vapor turbines, wherein the evaporatorportion first stage working fluid conduit is in communication with afirst vapor turbine and exhausts to the fourth stage working fluidconduit of the condenser portion.

In yet another aspect, first, second, third and fourth stage workingfluid conduits of the evaporator portion are in communication withfirst, second, third and fourth vapor turbines, wherein the evaporatorportion second stage working fluid conduit is in communication with asecond vapor turbine and exhausts to the third stage working fluidconduit of the condenser portion.

In yet another aspect, first, second, third and fourth stage workingfluid conduits of the evaporator portion are in communication withfirst, second, third and fourth vapor turbines, wherein the evaporatorportion third stage working fluid conduit is in communication with athird vapor turbine and exhausts to the second stage working fluidconduit of the condenser portion.

In yet another aspect, first, second, third and fourth stage workingfluid conduits of the evaporator portion are in communication withfirst, second, third and fourth vapor turbines, wherein the evaporatorportion fourth stage working fluid conduit is in communication with afourth vapor turbine and exhausts to the first stage working fluidconduit of the condenser portion.

In still a further aspect, a first electrical generator is driven by thefirst turbine, the fourth turbine, or a combination of the first andfourth turbine.

In still a further aspect, a second electrical generator is driven bythe second turbine, the third turbine, or a combination of both thesecond and third turbine.

Additional aspects can incorporate one or more of the followingfeatures: the first and fourth turbines or the second and third turbinesproduce between 9 MW and 60 MW of electrical power; the first and secondturbines produce approximately 55 MW of electrical power; the first andsecond turbines form one of a plurality of turbine-generator sets in anOcean Thermal Energy Conversion power plant; the first stage warm waterintake is free of interference from the second stage cold waterdischarge; the first stage cold water intake is free of interferencefrom the second stage warm water discharge; the working fluid within thefirst or second stage working fluid passages comprises a commercialrefrigerant. The working fluid comprises any fluid with suitablethermodynamic properties such as ammonia, propylene, butane, R-134, orR-22; the working fluid in the first and second stage working fluidpassages increases in temperature between 12° F. and 24° F.; a firstworking fluid flows through the first stage working fluid passage and asecond working fluid flows through the second stage working fluidpassage, wherein the second working fluid enters the second vaporturbine at a lower temperature than the first working fluid enters thefirst vapor turbine; the working fluid in the first and second stageworking fluid passages decreases in temperature between 12° F. and 24°F.; a first working fluid flows through the first stage working fluidpassage and a second working fluid flows through the second stageworking fluid passage, wherein the second working fluid enters thesecond deck portion at a lower temperature than the first working fluidenters the second deck portion.

Further aspects can also incorporate one or more of the followingfeatures: the warm water flowing within the first or second stage warmwater structural passage comprises, warm sea water, geo-thermally heatedwater, solar heated reservoir water; heated industrial cooling water, ora combination thereof; the warm water flows between 500,000 and6,000,000 gpm; the warm water flows at 5,440,000 gpm; the warm waterflows between 300,000,000 lb/hr and 1,000,000,000 lb/hr; the warm waterflows at 2,720,000 lb/hr; the cold water flowing within the first orsecond stage cold water structural passage comprises cold sea water,cold fresh water, cold subterranean water or a combination thereof; thecold water flows between 250,000 and 3,000,000 gpm; the cold water flowsat 3,420,000 gpm; the cold water flows between 125,000,000 lb/hr and1,750,000,000 lb/hr; the cold water flows at 1,710,000 lb/hr.

Aspects can also incorporate one or more of the following features: theoffshore structure is a low heave structure; the offshore structure is afloating spar structure; the offshore structure is a semi-submersiblestructure.

A still further aspect can include a high-volume, low-velocity heatexchange system for use in an ocean thermal energy conversion powerplant, comprising: a first stage cabinet that further comprises a firstwater flow passage for heat exchange with a working fluid; and a firstworking fluid passage; and a second stage cabinet coupled to the firststage cabinet, that further comprises a second water flow passage forheat exchange with a working fluid and coupled to the first water flowpassage in a manner to limit pressure drop of water flowing from thefirst water flow passage to the second water flow passage; and a secondworking fluid passage. The first and second stage cabinets comprisestructural members of the power plant.

In one aspect, water flows from the first stage cabinet to the secondstage cabinet and the second stage cabinet is beneath the first stagecabinet evaporator. In another aspect, water flows from the first stagecabinet to the second stage cabinet and the second stage cabinet isabove the first stage cabinet in the condensers and below the firststage cabinet in the evaporators.

In still a further aspect, a cold water pipe provides cold water fromocean depths to the cold water intake of the OTEC. The cold water intakecan be in the second deck portion of the submerged portion of the OTECplant. The cold water pipe can be a segmented construction. The coldwater pipe can be a continuous pipe. The cold water pipe can comprise:an elongated tubular structure having an outer surface, a top end and abottom end. The tubular structure can further comprise a plurality offirst and second stave segments wherein each stave segment has a topportion and a bottom portion, and wherein the top portion of the secondstave segment is offset from the top portion of the first stavedsegment. The cold water pipe can include a strake or ribbon at leastpartially wound spirally about the outer surface. The first and secondstaves and/or the strake can comprise polyvinyl chloride (PVC),chlorinated polyvinyl chloride (CPVC), fiber reinforced plastic (FRP),reinforced polymer mortar (RPMP), polypropylene (PP), polyethylene (PE),cross-linked high-density polyethylene (PEX), polybutylene (PB),acrylonitrile butadiene styrene (ABS); polyester, fiber reinforcedpolyester, vinyl ester, reinforced vinyl ester, concrete, ceramic, or acomposite of one or more thereof.

Further aspects include a dynamic connection between the submergedportion of the OTEC plant and the cold water pipe. The dynamicconnection can support the weight and dynamic forces of the cold waterpipe while it is suspended from the OTEC platform. The dynamic pipeconnection can allow for relative movement between the OTEC platform andthe cold water pipe. The relative movement can be between 0.5° and 30°from vertical. In one aspect the relative movement can be between 0.5°and 5° from vertical. The dynamic pipe connection can include aspherical or arcuate bearing surface.

In some embodiments, a static connection is provided between thesubmerged portion of the OTEC plant and the cold water pipe. In thesesystems, the top of the cold water pipe can be conical and is retractedinto a conical receptacle using lines and winches lowered from withinthe spar. The old water pipe can be retained using locking mechanismssuch that the lines can be detached for use in lifting equipment fromthe lower decks of the spar to the mid-body decks.

In an aspect, a submerged vertical pipe connection comprises a floatingstructure having a vertical pipe receiving bay, wherein the receivingbay has a first diameter, a vertical pipe for insertion into the pipereceiving bay, the vertical pipe having a second diameter smaller thanthe first diameter of the pipe receiving bay; a bearing surface; and oneor more detents operable with the bearing surface, wherein the detentsdefine a diameter that is different than the first or second diameterwhen in contact with the bearing surface.

More details of other aspects are described in U.S. patent applicationSer. No. ______ (Attorney Docket No. 25667-016001) entitled Staved OceanThermal Energy Conversion Power Plant—Cold Water Pipe Connection, andU.S. patent application Ser. No. ______ (Attorney Docket No.25667-009001) entitled Ocean Thermal Energy Conversion Power Plant,filed simultaneously with the present application and incorporatedherein by reference in their entirety.

Aspects may have one or more of the following advantages: OTEC powerproduction requires little to no fuel costs for energy production; thelow pressures and low temperatures involved in the OTEC heat enginereduce component costs and require ordinary materials compared to thehigh-cost, exotic materials used in high pressure, high temperaturepower generation plants; plant reliability is comparable to commercialrefrigeration systems, operating continuously for several years withoutsignificant maintenance; reduced construction times compared to highpressure, high temperature plants; and safe, environmentally benignoperation and power production. Additional advantages may include,increased net efficiency compared to traditional OTEC systems, lowersacrificial electrical loads; reduced pressure loss in warm and coldwater passages as well as working fluid flow passages; modularcomponents; less frequent off-grid production time; low heave andreduced susceptibility to wave action; discharge of cooling water belowsurface levels, intake of warm water free from interference from coldwater discharge.

The details of one or more embodiments are set forth in the accompanyingdrawings and the description below. Other features, objects, andadvantages will be apparent from the description and drawings, and fromthe claims.

DESCRIPTION OF DRAWINGS

FIG. 1 illustrates an exemplary prior-art OTEC heat engine.

FIG. 2 illustrates an exemplary prior-art OTEC power plant.

FIG. 3 illustrates OTEC structure.

FIG. 4 illustrates a deck plan of a heat exchanger deck.

FIG. 5 illustrates a cabinet heat exchanger.

FIG. 6A illustrates a conventional heat exchange cycle.

FIG. 6B illustrates a cascading multi-stage heat exchange cycle.

FIG. 6C illustrates a hybrid cascading multi-stage heat exchange cycle.

FIG. 6D illustrates the evaporator pressure drop and associate powerproduction.

FIGS. 7A and B illustrate an exemplary OTEC heat engine.

FIG. 8 illustrates a conventional shell and tube heat exchanger.

FIG. 9 illustrates a conventional plate heat exchanger.

FIG. 10 illustrates a cabinet heat exchanger.

FIG. 11 illustrates a perspective view of a heat exchange platearrangement.

FIG. 12 illustrates a perspective view of a heat exchange platearrangement.

FIG. 13 illustrates a side view of a heat exchange plate configuration.

FIG. 14 illustrates a P-h diagram of a conventional high temperaturesteam cycle.

FIG. 15 illustrates a P-h diagram of a heat cycle.

FIG. 16 illustrates an embodiment of a heat exchange plate.

FIG. 17 illustrates an embodiment of a heat exchange plate.

FIG. 18 illustrates a portion of a heat exchange plate.

FIGS. 19A and 19B illustrates an embodiment of a pair of heat exchangeplates.

FIGS. 20A and 20B illustrates an embodiment of a pair of heat exchangeplates.

FIGS. 21A-21D illustrate exemplary layout of heat exchange plates for anOTEC plant.

FIGS. 22A and 22B are, respectively, a schematic plan view of a heatexchange cassette and a cross-section of a working fluid channel of theheat exchange plate.

FIGS. 23A and 23B are, respectively, a schematic plan view of a heatexchange cassette and a cross-section of a working fluid channel of theheat exchange plate.

FIGS. 24A and 24B are, respectively, a schematic plan view of a heatexchange cassette and a cross-section of a working fluid channel of theheat exchange plate.

FIGS. 25A and 25B are, respectively, a schematic plan view of a heatexchange cassette and a cross-section of a working fluid channel of theheat exchange plate.

FIGS. 26A and 26B are, respectively, a schematic plan view of a heatexchange cassette and a cross-section of a working fluid channel of theheat exchange plate.

FIGS. 27A and 27B compare a full-size heat exchange cassette and ascaled-down heat exchange cassette.

FIGS. 28-32 are schematic plan views of embodiments of heat exchangecassettes.

FIGS. 33-35 illustrate cross-sections of working fluid channels.

Like reference symbols in the various drawings indicate like elements.

DETAILED DESCRIPTION

This disclosure relates to electrical power generation using OceanThermal Energy Conversion (OTEC) technology. Aspects relate to afloating OTEC power plant having improved overall efficiencies withreduced parasitic loads, greater stability, and lower construction andoperating costs over convention OTEC power plants. Other aspects includelarge volume water conduits that are integral with the floatingstructure. Modularity and compartmentation of the multi-stage OTEC heatengine reduces construction and maintenance costs, limits off-gridoperation and improves operating performance. Still further aspectsprovide for a floating platform having integrated heat exchangecompartments and provides for low movement of the platform due to waveaction. The integrated floating platform may also provide for efficientflow of the warm water or cool water through the multi-stage heatexchanger, increasing efficiency and reducing the parasitic powerdemand. In particular, highly efficient heat exchange plates can provideincreased overall efficiency, thus further reducing parasitic powerdemand. Aspects promote a neutral thermal footprint by discharging warmand cold water at appropriate depth/temperature ranges. Energy extractedin the form of electricity reduces the bulk temperature to the ocean.

OTEC is a process that uses heat energy from the sun that is stored inthe Earth's oceans to generate electricity. OTEC uses the temperaturedifference between the warmer, top layer of the ocean and the colder,deep ocean water. Typically this difference is at least 36° F. (20° C.).These conditions exist in tropical areas, roughly between the Tropic ofCapricorn and the Tropic of Cancer, or even 20° north and southlatitude. The OTEC process uses the temperature difference to power aRankine cycle, with the warm surface water serving as the heat sourceand the cold deep water serving as the heat sink. Rankine cycle turbinesdrive generators which produce electrical power.

FIG. 1 illustrates a typical OTEC Rankine cycle heat engine 10 whichincludes warm sea water inlet 12, evaporator 14, warm sea water outlet15, turbine 16, cold sea water inlet 18, condenser 20, cold sea wateroutlet 21, working fluid conduit 22 and working fluid pump 24.

In operation, heat engine 10 can use any one of a number of workingfluids, for example commercial refrigerants such as ammonia. Otherworking fluids can include propylene, butane, R-22 and R-134a. Othercommercial refrigerants can be used. Warm sea water betweenapproximately 75° F. and 85° F., or more, is drawn from the oceansurface or just below the ocean surface through warm sea water inlet 12and in turn warms the ammonia working fluid passing through evaporator14. The ammonia boils to a vapor pressure of approximately 9.3 atm. Thevapor is carried along working fluid conduit 22 to turbine 16. Theammonia vapor expands as it passes through the turbine 16, producingpower to drive an electric generator 25. The ammonia vapor then enterscondenser 20 where it is cooled to a liquid by cold sea water drawn froma deep ocean depth of approximately 3000 ft. The cold sea water entersthe condenser at a temperature of approximately 40° F. The vaporpressure of the ammonia working fluid at the temperature in thecondenser 20, approximately 51° F., is 6.1 atm. Thus, a significantpressure difference is available to drive the turbine 16 and generateelectric power. As the ammonia working fluid condenses, the liquidworking fluid is pumped back into the evaporator 14 by working fluidpump 24 via working fluid conduit 22.

The heat engine 10 of FIG. 1 is essentially the same as the Rankinecycle of most steam turbines, except that OTEC differs by usingdifferent working fluids and lower temperatures and pressures. The heatengine 10 of the FIG. 1 is also similar to commercial refrigerationplants, except that the OTEC cycle is run in the opposite direction sothat a heat source (e.g., warm ocean water) and a cold heat sink (e.g.,deep ocean water) are used to produce electric power.

FIG. 2 illustrates the components of a floating OTEC power plant 200,which include: the vessel or platform 210, warm sea water inlet 212,warm water pump 213, evaporator 214, warm sea water outlet 215,turbine-generator 216, cold water pipe 217, cold water inlet 218, coldwater pump 219, condenser 220, cold water outlet 221, working fluidconduit 222, working fluid pump 224, and pipe connections 230. OTECplant 200 can also include electrical generation, transformation andtransmission systems, position control systems such as propulsion,thrusters, or mooring systems, as well as various auxiliary and supportsystems (for example, personnel accommodations, emergency power, potablewater, black and grey water, firefighting, damage control, reservebuoyancy, and other common shipboard or marine systems).

Implementations of OTEC power plants utilizing the basic heat engine andsystem of FIGS. 1 and 2 have a relatively low overall efficiency of 3%or below. Because of this low thermal efficiency, OTEC operationsrequire the flow of large amounts of water through the power system perkilowatt of power generated. This in turn requires large heat exchangershaving large heat exchange surface areas.

Such large volumes of water and large surface areas require considerablepumping capacity in the warm water pump 213 and cold water pump 219,reducing the net electrical power available for distribution to ashore-based facility or on board industrial purposes. Moreover, thelimited space of most surface vessels does not easily facilitate largevolumes of water directed to and flowing through the evaporator orcondenser. Indeed, large volumes of water require large diameter pipesand conduits. Putting such structures in limited space requires multiplebends to accommodate other machinery. And the limited space of typicalsurface vessels or structures does not easily facilitate the large heatexchange surface area required for maximum efficiency in an OTEC plant.Thus the OTEC systems and vessel or platform have traditionally beenlarge and costly. This has led to an industry conclusion that OTECoperations are a high cost, low yield energy production option whencompared to other energy production options using higher temperaturesand pressures.

The systems and approaches described herein address technical challengesin order to improve the efficiency of OTEC operations and reduce thecost of construction and operation.

The vessel or platform 210 requires low motions to limit dynamic forcesbetween the cold water pipe 217 and the vessel or platform 210 and toprovide a benign operating environment for the OTEC equipment in theplatform or vessel. The vessel or platform 210 should also support coldand warm water inlet (218 and 212) volume flows, bringing in sufficientcold and warm water at appropriate levels to ensure OTEC processefficiency. The vessel or platform 210 should also enable cold and warmwater discharge via cold and warm water outlets (221 and 215) well belowthe waterline of vessel or platform 210 to avoid thermal recirculationinto the ocean surface layer. Additionally, the vessel or platform 210should survive heavy weather without disrupting power generatingoperations.

The OTEC heat engine 10 described herein uses a highly efficient thermalcycle for maximum efficiency and power production. Heat transfer inboiling and condensing processes, as well as the heat exchangermaterials and design, limit the amount of energy that can be extractedfrom each pound of warm seawater. The heat exchangers used in theevaporator 214 and the condenser 220 use high volumes of warm and coldwater flow with low head loss to limit parasitic loads. The heatexchangers also provide high coefficients of heat transfer to enhanceefficiency. The heat exchangers incorporate materials and designstailored to the warm and cold water inlet temperatures to enhanceefficiency. The heat exchanger design can use a simple constructionmethod with low amounts of material to reduce cost and volume.

The turbo generators 216 are highly efficient with low internal lossesand may also be tailored to the working fluid to enhance efficiency

FIG. 3 illustrates an implementation of an OTEC system that enhances theefficiency of previous OTEC power plants and overcomes many of thetechnical challenges associated therewith. This implementation comprisesa spar for the vessel or platform, with heat exchangers and associatedwarm and cold water piping integral to the spar.

OTEC Spar 310 houses an integral multi-stage heat exchange system foruse with an OTEC power generation plant. Spar 310 includes a submergedportion 311 below waterline 305. Submerged portion 311 comprises warmwater intake portion 340, evaporator portion 344, warm water dischargeportion 346, condenser portion 348, cold water intake portion 350, coldwater pipe 351, cold water discharge portion 352, machinery deck portion354, and deck house 360.

In operation, warm sea water of between 75° F. and 85° F. is drawnthrough warm water intake portion 340 and flows down the spar thoughstructurally integral warm water conduits (not shown). Due to the highvolume water flow requirements of OTEC heat engines, the warm waterconduits direct flow to the evaporator portion 344 of between 500,000gpm and 6,000,000 gpm. The warm water conduits have a diameter ofbetween 6 ft and 35 ft, or more. Due to this size, the warm waterconduits are vertical structural members of spar 310. Warm waterconduits can be large diameter pipes of sufficient strength tovertically support spar 310. Alternatively, the warm water conduits canbe passages integral to the construction of the spar 310.

Warm water then flows through the evaporator portion 344 which housesone or more stacked, multi-stage heat exchangers for warming a workingfluid to a vapor. The warm sea water is then discharged from spar 310via warm water discharge 346. Warm water discharge can be located ordirected via a warm water discharge pipe to a depth at or close to anocean thermal layer that is approximately the same temperature as thewarm water discharge temperature to limit environmental impacts. Thewarm water discharge can be directed to a sufficient depth to avoidthermal recirculation with either the warm water intake or cold waterintake.

Cold sea water is drawn from a depth of between 2500 and 4200 ft, ormore, at a temperature of approximately 40° F., via cold water pipe 351.The cold sea water enters spar 310 via cold water intake portion 350.Due to the high volume water flow requirements of OTEC heat engines, thecold sea water conduits direct flow to the condenser portion 348 ofbetween 500,000 gpm and 3,500,000 gpm. Such cold sea water conduits havea diameter of between 6 ft and 35 ft, or more. Due to this size, thecold sea water conduits are vertical structural members of spar 310.Cold water conduits can be large diameter pipes of sufficient strengthto vertically support spar 310. Alternatively, the cold water conduitscan be passages integral to the construction of the spar 310.

Cold sea water then flows upward to stacked multi-stage condenserportion 348, where the cold sea water cools a working fluid to a liquid.The cold sea water is then discharged from spar 310 via cold sea waterdischarge 352. Cold water discharge can be located or directed via acold sea water discharge pipe to depth at or close to an ocean thermallayer that is approximately the same temperature as the cold sea waterdischarge temperature. The cold water discharge can be directed to asufficient depth to avoid thermal recirculation with either the warmwater intake or cold water intake.

Machinery deck portion 354 can be positioned vertically between theevaporator portion 344 and the condenser portion 348. Positioningmachinery deck portion 354 beneath evaporator portion 344 allows nearlystraight line warm water flow from intake, through the multi-stageevaporators, and to discharge. Positioning machinery deck portion 354above condenser portion 348 allows nearly straight line cold water flowfrom intake, through the multi-stage condensers, and to discharge.Machinery deck portion 354 includes turbo generators 356. In operation,warm working fluid heated to a vapor flows from evaporator portion 344to one or more turbo generators 356. The working fluid expands in turbogenerator 356 thereby driving a turbine for the production of electricalpower. The working fluid then flows to condenser portion 348 where it iscooled to a liquid and pumped to evaporator portion 344.

FIG. 4 illustrates an implementation of an OTEC system wherein aplurality of multi-stage heat exchangers 420 is arranged about theperiphery of OTEC spar 410. Heat exchangers 420 can be evaporators orcondensers used in an OTEC heat engine. The peripheral layout of heatexchanges can be used with evaporator portion 344 or condenser portion348 of an OTEC spar platform (as shown in FIG. 3). The peripheralarrangement can support any number of heat exchangers (e.g., 1 heatexchanger, between 2 and 8 heat exchangers, 8-16 heat exchanger, 16-32heat exchangers, or 32 or more heat exchangers). One or more heatexchangers can be peripherally arranged on a single deck or on multipledecks (e.g., on 2, 3, 4, 5, or 6 or more decks) of the OTEC spar 410.One or more heat exchangers can be peripherally offset between two ormore decks such that no two heat exchangers are vertically aligned overone another. One or more heat exchangers can be peripherally arranged sothat heat exchangers in one deck are vertically aligned with heatexchanges on another adjacent deck.

Individual heat exchangers 420 can comprise a multi-stage heat exchangesystem (e.g., a 2, 3, 4, 5, or 6 or more heat exchange system). In someembodiments, individual heat exchangers 420 are cabinet heat exchangersconstructed to provide low pressure loss in the warm sea water flow,cold sea water flow, and working fluid flow through the heat exchanger.

Referring to FIG. 5, an embodiment of a cabinet heat exchanger 520includes multiple heat exchange stages, 521, 522, 523 and 524. In someimplementations, the stacked heat exchangers accommodate warm sea waterflowing down through the cabinet, from first evaporator stage 521, tosecond evaporator stage 522, to third evaporator stage 523 to fourthevaporator stage 524. In another embodiment of the stacked heat exchangecabinet, cold sea water flows up through the cabinet from firstcondenser stage 531, to second condenser stage 532, to third condenserstage 533, to fourth condenser stage 534. Working fluid flows throughworking fluid supply conduits 538 and working fluid discharge conduits539. In an embodiment, working fluid conduits 538 and 539 enter and exiteach heat exchanger stage horizontally as compared to the vertical flowof the warm sea water or cold sea water. The vertical multi-stage heatexchange design of cabinet heat exchanger 520 facilitates an integratedvessel (e.g., spar) and heat exchanger design, removes the requirementfor interconnecting piping between heat exchanger stages, and ensuresthat virtually all of the heat exchanger system pressure drop occursover the heat transfer surface.

The heat transfer surface efficiency can be improved using surfaceshape, treatment and spacing as described herein. Material selectionsuch as alloys of aluminum offer superior economic performance overtraditional titanium base designs. The heat transfer surface cancomprise 100 Series, 3000 Series, or 5000 Series aluminum alloys. Theheat transfer surface can comprise titanium and titanium alloys.

It has been found that the multi-stage heat exchanger cabinet enableshigh energy transfer to the working fluid from the sea water within therelatively low available temperature differential of the OTEC heatengine. The thermodynamic efficiency of any OTEC power plant is afunction of how close the temperature of the working fluid approachesthat of the sea water. The physics of the heat transfer dictate that thearea required to transfer the energy increases as the temperature of theworking fluid approaches that of the sea water. Increasing the velocityof the sea water can increase the heat transfer coefficient to offsetthe increase in surface area. However, increasing the velocity of thesea water can greatly increases the power required for pumping, therebyincreasing the parasitic electrical load on the OTEC plant.

FIG. 6A illustrates an OTEC cycle wherein the working fluid is boiled ina heat exchanger using warm surface sea water. The fluid properties inthis conventional Rankine cycle are constrained by the boiling processthat limits the leaving working fluid to approximately 3° F. below theleaving warm seawater temperature. In a similar fashion, the condensingside of the cycle is limited to being no close than 2° F. higher thanthe leaving cold seawater temperature. The total available temperaturedrop for the working fluid is approximately 12° F. (between 68° F. and56° F.).

It has been found that a cascading multi-stage OTEC cycle allows theworking fluid temperatures to more closely match that of the sea water.This increase in temperature differential increases the amount of workthat can be done by the turbines associated with the OTEC heat engine.

FIG. 6B illustrates a cascading multi-stage OTEC cycle using multiplesteps of boiling and condensing to expand the available working fluidtemperature drop. Each step requires an independent heat exchanger, or adedicated heat exchanger stage in the cabinet heat exchanger 520 of FIG.5. The cascading multi-stage OTEC cycle of FIG. 6 b allows for matchingthe output of the turbines with the expected pumping loads for the seawater and working fluid. This highly efficient design would requirededicated and customized turbines.

FIG. 6C illustrates a hybrid yet still efficient cascading OTEC cyclethat facilitates the use of identical equipment (e.g., turbines,generators, pumps) while retaining the thermodynamic efficiencies oroptimization of the true cascade arrangement of FIG. 6B. In the hybridcascade cycle of FIG. 6C, the available temperature differential for theworking fluid ranges from about 18° F. to about 22° F. This narrow rangeallows the turbines in the heat engine to have identical performancespecifications, thereby lowering construction and operation costs.

System performance and power output is greatly increased using thehybrid cascade cycle in an OTEC power plant. Table A compares theperformance of the conventional cycle of FIG. 6A with that of the hybridcascading cycle of FIG. 6C.

TABLE A Estimated Performance for 100 MW Net Output Four Stage HybridConventional Cycle Cascade Cycle Warm Sea Water 4,800,000 GPM 3,800,000GPM Flow Cold Sea Water 3,520,000 GPM 2,280,000 GPM Flow Gross Heat Rate163,000 BTU/kWH 110,500 BTU/kWHUtilizing the four stage hybrid cascade heat exchange cycle reduces theamount of energy that needs to be transferred between the fluids. Thisin turn serves to reduce the amount of heat exchange surface that isrequired.

The performance of heat exchangers is affected by the availabletemperature difference between the fluids as well as the heat transfercoefficient at the surfaces of the heat exchanger. The heat transfercoefficient generally varies with the velocity of the fluid across theheat transfer surfaces. Higher fluid velocities require higher pumpingpower, thereby reducing the net efficiency of the plant. A hybridcascading multi-stage heat exchange system facilitates lower fluidvelocities and greater plant efficiencies. The stacked hybrid cascadeheat exchange design also facilitates lower pressure drops through theheat exchanger. And the vertical plant design facilitates lower pressuredrop across the whole system.

FIG. 6D illustrates the impact of heat exchanger pressure drop on thetotal OTEC plant generation to deliver 100 MW to a power grid. Limitingpressure drop through the heat exchanger greatly enhances the OTEC powerplant's performance. Pressure drop is reduced by providing an integratedvessel or platform-heat exchanger system, wherein the sea water conduitsform structural members of the vessel and allow for sea water flow fromone heat exchanger stage to another in series. An approximate straightline sea water flow, with low changes in direction from intake into thevessel, through the pump, through the heat exchange cabinets and in turnthrough each heat exchange stage in series, and ultimate dischargingfrom the plant, allows for low pressure drop.

Cascade Hybrid OTEC Power Generation:

An integrated multi-stage OTEC power plant can produce electricity usingthe temperature differential between the surface water and deep oceanwater in tropical and subtropical regions. Traditional piping runs forsea water can be eliminated by using the off-shore vessel's orplatform's structure as a conduit or flow passage. Alternatively, thewarm and cold sea water piping runs can use conduits or pipes ofsufficient size and strength to provide vertical or other structuralsupport to the vessel or platform. These integral sea water conduitsections or passages serve as structural members of the vessel, therebyreducing the requirements for additional steel. As part of the integralsea water passages, multi-stage cabinet heat exchangers providesmultiple stages of working fluid evaporation without the need forexternal water nozzles or piping connections. The integrated multi-stageOTEC power plant allows the warm and cold sea water to flow in theirnatural directions. The warm sea water flows downward through the vesselas it is cooled before being discharged into a cooler zone of the ocean.In a similar fashion, the cold sea water from deep in the ocean flowsupward through the vessel as it is warmed before discharging into awarmer zone of the ocean. This arrangement avoids the need for changesin sea water flow direction and associated pressure losses. Thearrangement also reduces the pumping energy required.

Multi-stage cabinet heat exchangers allow for the use of a hybridcascade OTEC cycle. These stacks of heat exchangers comprise multipleheat exchanger stages or sections that have sea water passing throughthem in series to boil or condense the working fluid as appropriate. Inthe evaporator section, the warm sea water passes through a first stagewhere it boils off some of the working fluid as the sea water is cooled.The warm sea water then flows down the stack into the next heatexchanger stage and boils off additional working fluid at a slightlylower pressure and temperature. This occurs sequentially through theentire stack. Each stage or section of the cabinet heat exchangersupplies working fluid vapor to a dedicated turbine that generateselectrical power. Each of the evaporator stages has a correspondingcondenser stage at the exhaust of the turbine. The cold sea water passesthrough the condenser stacks in a reverse order to the evaporators.

Referring to FIGS. 7A and 7B, an exemplary multi-stage OTEC heat engine710 utilizes a hybrid cascading heat exchange cycles. Warm sea water ispumped from a warm sea water intake (not shown) by warm water pump 712,discharging from the pump at approximately 1,360,000 gpm and at atemperature of approximately 79° F. All or parts of the warm waterconduit from the warm water intake to the warm water pump, and from thewarm water pump to the stacked heat exchanger cabinet can form integralstructural members of the vessel.

From the warm water pump 712, the warm sea water then enters a firststage evaporator 714 where it boils a first working fluid. The warmwater exits the first stage evaporator 714 at a temperature ofapproximately 76.8° F. and flows down to a second stage evaporator 715.

The warm water enters the second stage evaporator 715 at approximately76.8° F. where it boils a second working fluid and exits the secondstage evaporator 715 at a temperature of approximately 74.5°.

The warm water flows down to a third stage evaporator 716 from thesecond stage evaporator 715, entering at a temperature of approximately74.5° F., where it boils a third working fluid. The warm water exits thethird stage evaporator 716 at a temperature of approximately 72.3° F.

The warm water then flows from the third stage evaporator 716 down tothe fourth stage evaporator 717, entering at a temperature ofapproximately 72.3° F., where it boils a fourth working fluid. The warmwater exits the fourth stage evaporator 717 at a temperature ofapproximately 70.1° F. and then discharges from the vessel. Though notshown, the discharge can be directed to a thermal layer at an oceandepth of approximately the same temperature as the discharge temperatureof the warm sea water. Alternately, the portion of the power planthousing the multi-stage evaporator can be located at a depth within thestructure so that the warm water is discharged to an appropriate oceanthermal layer. In some embodiments, the warm water conduit from thefourth stage evaporator to the warm water discharge of the vessel cancomprise structural members of the vessel.

Similarly, cold sea water is pumped from a cold sea water intake (notshown) via cold sea water pump 722, discharging from the pump atapproximately 855,003 gpm and at a temperature of approximately 40.0° F.The cold sea water is drawn from ocean depths of between approximately2700 and 4200 ft, or more. The cold water conduit carrying cold seawater from the cold water intake of the vessel to the cold water pump,and from the cold water pump to the first stage condenser can comprisein its entirety or in part structural members of the vessel.

From cold sea water pump 722, the cold sea water enters a first stagecondenser 724, where it condenses the fourth working fluid from thefourth stage boiler 717. The cold seawater exits the first stagecondenser at a temperature of approximately 43.5° F. and flows up to asecond stage condenser 725.

The cold sea water enters the second stage condenser 725 atapproximately 43.5° F. where it condenses the third working fluid fromthird stage evaporator 716. The cold sea water exits the second stagecondenser 725 at a temperature approximately 46.9° F. and flows up to athird stage condenser 726.

The cold sea water enters the third stage condenser 726 at a temperatureof approximately 46.9° F. where it condenses the second working fluidfrom second stage evaporator 715. The cold sea water exits the thirdstage condenser 726 at a temperature approximately 50.4° F.

The cold sea water then flows up from the third stage condenser 726 to afourth stage condenser 727, entering at a temperature of approximately50.4° F. In the fourth stage condenser, the cold sea water condenses thefirst working fluid from the first stage evaporator 714. The cold seawater then exits the fourth stage condenser at a temperature ofapproximately 54.0° F. and ultimately discharges from the vessel. Thecold sea water discharge can be directed to a thermal layer at an oceandepth of or approximately the same temperature as the dischargetemperature of the cold sea water. Alternately, the portion of the powerplant housing the multi-stage condenser can be located at a depth withinthe structure so that the cold sea water is discharged to an appropriateocean thermal layer.

The first working fluid enters the first stage evaporator 714 at atemperature of 56.7° F. where it is heated to a vapor with a temperatureof 74.7° F. The first working fluid then flows to first turbine 731 andthen to the fourth stage condenser 727 where the first working fluid iscondensed to a liquid with a temperature of approximately 56.5° F. Theliquid first working fluid is then pumped via first working fluid pump741 back to the first stage evaporator 714.

The second working fluid enters the second stage evaporator 715 at atemperature approximately 53.0° F. where it is heated to a vapor. Thesecond working fluid exits the second stage evaporator 715 at atemperature approximately 72.4° F. The second working fluid then flow toa second turbine 732 and then to the third stage condenser 726. Thesecond working fluid exits the third stage condenser at a temperatureapproximately 53.0° F. and flows to working fluid pump 742, which inturn pumps the second working fluid back to the second stage evaporator715.

The third working fluid enters the third stage evaporator 716 at atemperature approximately 49.5° F. where it will be heated to a vaporand exit the third stage evaporator 716 at a temperature ofapproximately 70.2° F. The third working fluid then flows to thirdturbine 733 and then to the second stage condenser 725 where the thirdworking fluid is condensed to a fluid at a temperature approximately49.5° F. The third working fluid exits the second stage condenser 725and is pumped back to the third stage evaporator 716 via third workingfluid pump 743.

The fourth working fluid enters the fourth stage evaporator 717 at atemperature of approximately 46.0° F. where it will be heated to avapor. The fourth working fluid exits the fourth stage evaporator 717 ata temperature approximately 68.0° F. and flow to a fourth turbine 734.The fourth working fluid exits fourth turbine 734 and flows to the firststage condenser 724 where it is condensed to a liquid with a temperatureapproximately 46.0° F. The fourth working fluid exits the first stagecondenser 724 and is pumped back to the fourth stage evaporator 717 viafourth working fluid pump 744.

The first turbine 731 and the fourth turbine 734 cooperatively drive afirst generator 751 and form first turbo generator pair 761. First turbogenerator pair will produce approximately 25 MW of electric power.

The second turbine 732 and the third turbine 733 cooperatively drive asecond generator 752 and form second turbo generator pair 762. Secondturbo generator pair 762 will produce approximately 25 MW of electricpower.

The four stage hybrid cascade heat exchange cycle of FIG. 7 allows themaximum amount of energy to be extracted from the relatively lowtemperature differential between the warm sea water and the cold seawater. Moreover, all heat exchangers can directly support turbogenerator pairs that produce electricity using the same componentturbines and generators.

It will be appreciated that multiple multi-stage hybrid cascading heatexchangers and turbo generator pairs can be incorporated into a vesselor platform design.

Multi-Stage, Open-Flow, Heat Exchange Cabinets

OTEC systems, by their nature require large volumes of water, forexample, a 100 megawatt OTEC power plant can require, for example, up toorders of magnitude more water than required for a similarly sizedcombustion fired steam power plant. In an exemplary implementation, a 25MW OTEC power plant can require approximately 1,000,000 gallons perminute of warm water supply to the evaporators and approximately 875,000gallons per minute of cold water to the condensers. The energy requiredfor pumping water together with the small temperature differentials(approximately 35 to 45 degrees F.) act to drive down efficiency whileraising the cost of construction.

Presently available heat exchangers are insufficient to handle the largevolumes of water and high efficiencies required for OTEC heat exchangeoperations. Shell and tube heat exchangers consist of a series of tubes.One set of these tubes contains the working fluid that must be eitherheated or cooled. The second non-working fluid runs over the tubes thatare being heated or cooled so that it can either provide the heat orabsorb the heat required. A set of tubes is called the tube bundle andcan be made up of several types of tubes: plain, longitudinally finned,etc. Shell and tube heat exchangers are typically used for high-pressureapplications. This is because the shell and tube heat exchangers arerobust due to their shape. Shell and tube exchangers are not ideal forthe low temperature differential, low pressure, high volume nature ofOTEC operations. For example, shell and tube heat exchangers, as shownin FIG. 8, typically require complicated piping arrangements with highpressure losses and associated piping energy. These types of heatexchangers are difficult to fabricate, install and maintain,particularly in a dynamic environment such as an offshore platform.Shell and tube heat exchanges also require precision assemblyparticularly for the shell to tube connections and for the internalsupports. Moreover, shell and tube heat exchangers often have a low heattransfer coefficient and are restricted in the volume of water that canbe accommodated.

FIG. 9 depicts a plate heat exchanger. Plate heat exchangers can includemultiple, thin, slightly-separated plates that have very large surfaceareas and fluid flow passages for heat transfer. This stacked-platearrangement can be more effective, in a given space, than the shell andtube heat exchanger. Advances in gasket and brazing technology have madethe plate-type heat exchanger increasingly practical. In HVACapplications for example, large heat exchangers of this type are calledplate-and-frame; when used in open loops, these heat exchangers arenormally of the gasket type to allow periodic disassembly, cleaning, andinspection. Permanently-bonded plate heat exchangers, such as dip-brazedand vacuum-brazed plate varieties, are often specified for closed-loopapplications such as refrigeration. Plate heat exchangers also differ inthe types of plates that are used, and in the configurations of thoseplates. Some plates may be stamped with “chevron” or other patterns,where others may have machined fins and/or grooves.

Plate heat exchangers, however, have some significant disadvantages inOTEC applications. For example, these types of heat exchangers canrequire complicated piping arrangements that do not easily accommodatethe large volumes of water needed with OTEC systems. Often, gaskets mustbe precisely fitted and maintained between each plate pair, andsignificant bolting is needed to maintain the gasket seals. Plate heatexchangers typically require complete disassembly to inspect and repaireven one faulty plate. Materials needed for plate heat exchangers arecan be limited to costly titanium and/or stainless steel. These types ofheat exchangers require relatively equal flow areas between the workingand non-working fluids. Flow ratios between the fluids are typically1:1. As can be seen in FIG. 9, supply and discharge ports are typicallyprovided on the face of the plate, reducing the total heat exchangesurface area and complicating the flow path of each of the working andnon-working fluids. Moreover, plate heat exchangers include complexinternal circuiting for nozzles that penetrate all plates.

In order to overcome the limitations of such conventional heatexchangers, a gasket-free, open flow heat exchanger is provided. In someimplementations, individual plates are horizontally aligned in a cabinetsuch that a gap exists between each plate. A flow path for the workingfluid runs through the interior of each plate in a pattern providinghigh heat transfer (e.g., alternating serpentine, chevrons, z-patterns,and the like). The working fluid enters each plate through connectionson the side of the plates so as to reduce obstructions in the face ofthe plate or impediments to the water flow by the working fluid. Thenon-working fluid, such as raw water, flows vertically through thecabinet and fills the gaps between each of the open-flow plates. In someimplementations, the non-working fluid is in contact with all sides ofthe open-flow plates or in contact with just the front and back surfacesof the open-flow plates.

FIG. 10 illustrates a stacked cabinet arrangement 520 of heatexchangers, similar to the arrangement as described in FIG. 5, with adetail of a single cabinet 524 having a rack of multiple heat exchangeplates 1022. The non-working fluid flows vertically through the cabinet524 and past each of the plates 1022 in the rack. Arrow 525 indicatesthe flow direction of the water. The flow direction of the water can befrom top to bottom or bottom to top. In some embodiments, the flowdirection can be in the natural direction of the water as it is heatedor cooled. For example, when condensing a working fluid, the water canflow through the cabinet arrangement from bottom to top in the naturalflow of convection as the water is warmed. In another example, whenevaporating a working fluid, the water can flow from top to bottom asthe water cools. In still other embodiments, the non-working fluid flowcan be horizontally across the cabinet, that is, from left to right orright to left.

Referring to FIG. 10, open-flow heat exchange cabinet 524 includescabinet face 1030 and cabinet side 1031. Opposite of cabinet face 1030is cabinet face 1032 (not shown) and opposite of cabinet side 1031 iscabinet side 1033 (not shown). The cabinet faces and sides form a plenumor water conduit through which the raw water non-working fluid flowswith little to no pressure losses due to piping. In contrast to thegasket heat exchanger described above with respect to FIG. 9, the openflow heat exchanger uses the cabinet to form a flow chamber containingthe non-working fluid (e.g., sea water) rather than using gasketsbetween plates to form the flow chamber containing the non-workingfluid. Thus, the open-flow heat exchange cabinet 524 is effectivelygasket-free. This aspect of this system provides significant advantagesover other plate and frame heat exchangers that rely on gaskets toisolate the working fluid from the energy providing medium (e.g., seawater). Corrosion testing of aluminum plate and frame heat exchangersdone at NELHA in the 1980s and 1990s had to stop after only six monthsbecause there was so much leakage around the gaskets where biologicaldeposits caused extensive erosion. The applicants identified gasketissues as a major impediment to using a plate and frame design in anOTEC system.

In addition, the cabinet approach combined with side mounted inlet andoutlet ports for the heat exchange plates avoids the needs for thesupply and discharge ports typically provided on the face of the plateheat exchange systems (see, e.g., FIG. 9). This approach increases thetotal heat exchange surface area of each plate as well as simplifyingthe flow path of both the working and non-working fluids. Removing thegaskets between the plates also removes significant obstructions thatcan cause resistance to flow. The gasket-free open-flow heat exchangecabinets can reduce back pressure and associated pumping demand, thusreducing the parasitic load of an OTEC plant and resulting in increasedpower that can be delivered to the utility company.

In the case of an OTEC condenser, cabinet 524 is open on the bottom tothe cold raw water supply, and open on the top to provide unobstructedfluid communication with the cabinet 523 above. The final cabinet in thevertical series 521 is open at the top to the raw water dischargesystem.

In the case of an evaporator, cabinet 521 is open on the top to the warmraw water supply and open at the bottom to provide unobstructed fluidcommunication to the cabinet below 522. The final cabinet 524 in thevertical series is open on the bottom to the warm raw water dischargesystem.

Within each of the heat exchange cabinets, a plurality of open-flow heatexchange plates 1022 are arranged in horizontal alignment to provide agap 1025 between each pair of plates 1022. Each open flow plate has afront face, a back face, a top surface, a bottom surface, and left andright sides. The plates 1022 are arranged in horizontal alignment sothat the back face of a first plate faces the front face of a secondplate immediately behind the first plate. A working fluid supply anddischarge are provided on the sides of each of the plates to avoidimpediments to the flow of the raw water through the gaps 1025 as theraw water flows past the front and back faces of the plurality of plates1022 in the rack. Each of the plates 1022 includes a working fluid flowpassage that is internal to the plate. Open-flow plates 1022 aredescribed in greater detail further below.

In some implementations, each individual plate 1022 has a dedicatedworking fluid supply and discharge such that the working fluid flowsthrough a single plate. Supply of the working fluid is directly to oneor more of working fluid supply passages. In other implementations, theworking fluid can flow through two or more plates in series before beingdischarged from the heat exchange cabinet to the reminder of the workingfluid system.

It will be appreciated that each heat exchanger cabinet 524, 523, 522,and 521 has similar components and is vertically aligned such that thehorizontally aligned plates 1022 in one cabinet vertically align overthe plates in the cabinet below. The gaps 1025 between plates 1022 onone cabinet vertically align over the gaps 1025 between plates 1022 inthe cabinet below.

Referring to FIGS. 11 and 12, an exemplary implementation of the platearrangement in heat exchange cabinet 524 includes a first open-flow heatexchange plate 1051 having an exterior surface including at least afront and back face. The exterior surface is in fluid communication withand surrounded by a non-working fluid 1057 such as cold raw water. Thefirst open flow plate also includes an internal passage in fluidcommunication with a working fluid 1058 flowing through the internalpassage. At least one more second open-flow heat exchange plate 1052 ishorizontally aligned with the first open-flow heat exchange plate 1051such that the front exterior surface of the second plate 1052 faces theback exterior surface of the first plate 1051. Like the first plate, theat least one second plate 1052 includes an exterior surface in fluidcommunication with and surrounded by the non-working fluid 1057, and aninternal passage in fluid communication with a working fluid 1058flowing through the internal passage. The first open-flow heat exchangeplate 1051 is separated from the second heat exchange plate 1052 by agap 1053. The non-working fluid 1057 flows through the gap.

FIG. 13 depicts a side view of an exemplary open-flow heat exchangecabinet 524 including a first open flow heat exchange plate 1051, asecond heat exchange plate 1052, and gaps 1053 separating each firstplate 1051 and 1052. The working fluid 1058 flows through the internalworking fluid flow passages 1055.

As described above, in some implementations, a single heat exchangecabinet can be dedicated to a single stage of a hybrid cascade OTECcycle. In some implementations, four heat exchange cabinets arevertically aligned, as depicted and described in FIG. 5. In otherimplementations, cabinets having working fluid supply and dischargelines connected to the sides of each plate can be used. This avoidsworking fluid conduits being on the face of the plates and impeding theflow of both the working fluid and the non-working fluid.

For example, a gasket-free multi-stage heat exchange system can includea first stage heat exchange rack comprising one or more open-flow platesin fluid communication with a first working fluid flowing through aninternal passage in each of the one or more open-flow plates. Theworking fluid can be supplied and discharged from each plate via supplyand discharge lines dedicate to each individual plate. A second stageheat exchange rack vertically aligned with the first heat exchange rackis also included. The second stage heat exchange rack comprising one ormore open-flow plates in fluid communication with a second working fluidflowing through an internal passage in each of the one or more open-flowplates. Again, the second working fluid is supplied and discharged toand from each individual plate through lines dedicated to eachindividual plate. A non-working fluid, such as raw water, flows firstthrough the first stage heat exchange rack and around each of the one ormore open-flow plates allowing for thermal exchange with the firstworking fluid. The non-working fluid then passes through the second heatexchange rack and around each of the open-flow plates allowing forthermal exchange with the second working fluid.

The first stage rack includes a plurality of open-flow plates inhorizontal alignment having a gap between each plate. The second stagerack also includes a plurality of open-flow plates in horizontalalignment having a gap between each plate within the second stage racks.The plurality of open-flow plates and gaps in the second stage rack arevertically aligned with the plurality of open-flow plates and gaps inthe first stage rack. This reduces pressure losses in the flow of thenon-working fluid through the first and second stage racks. Pressurelosses in the non-working fluid are also reduced by having thenon-working fluid directly discharge from one cabinet to the nextthereby eliminating the need for extensive and massive piping systems.In some embodiments, the walls of the cabinets containing the first andsecond stage racks of heat exchange plates form the conduit throughwhich the non-working fluid flows.

Due to the open-flow arrangement of the plates in each rack of eachstage of an exemplary four stage OTEC system, the flow ratio of thenon-working fluid to the working fluid is increased from the typical 1:1of most conventional plate heat exchanger systems. In someimplementations the flow ratio of the non-working fluid is greater than1:1, (e.g., greater than 2:1, greater than 10:1, greater than 20:1,greater than 30:1, greater than 40:1, greater than 50:1, greater than60:1, greater than 70:1, greater than 80:1, greater than 90:1 or greaterthan 100:1).

When a multi-stage arrangement of heat exchange cabinets is used as acondenser, the non-working fluid (e.g., the cold sea water) generallyenters the first stage cabinet at a temperature lower than when thenon-working fluid enters the second stage cabinet, and the non-workingfluid then enters the second stage cabinet at a temperature lower thanwhen the non-working fluid entered the third stage cabinet; and thenon-working fluid enters the third stage cabinet at a temperaturegenerally lower than when it enters the fourth stage cabinet.

When a multi-stage arrangement of heat exchange cabinets are used as anevaporator, the non-working fluid (e.g., the warm sea water) generallyenters the first stage cabinet at a temperature higher than when thenon-working fluid enters the second stage cabinet, and the non-workingfluid then enters the second stage cabinet at a temperature higher thanwhen the non-working fluid enters the third stage cabinet; and thenon-working fluid enters the third stage cabinet at a temperaturegenerally higher than when it enters the fourth stage cabinet.

When a multi-stage arrangement of heat exchange cabinets are used as ancondenser, the working fluid (e.g., the ammonia) generally exits thefirst stage cabinet a temperature lower than when the working fluidexits the second stage cabinet, and the working fluid exits the secondstage cabinet at a temperature lower than the working fluid exits thethird stage cabinet; and the working fluid exits the third stage cabinetat a temperature generally lower than when it exits the fourth stagecabinet.

When a multi-stage arrangement of heat exchange cabinets are used as anevaporator, the working fluid (e.g., the ammonia) generally exits thefirst stage cabinet at a temperature higher than the working fluidexiting the second stage cabinet, and the working fluid exits the secondstage cabinet at a temperature generally higher than the working fluidexits the third stage cabinet; and the working fluid exits the thirdstage cabinet at a temperature generally higher than when it exits thefourth stage cabinet.

An exemplary heat balance of an implementation of a four stage OTECcycle is described herein and generally illustrates these concepts.

In some implementations, a four stage, gasket-free, heat exchange systemincludes a first stage heat exchange rack having one or more open-flowplates, each plate includes an exterior surface having at least a frontand back face surrounded by a non-working fluid. Each plate alsoincludes an internal passage in fluid communication with a first workingfluid flowing through the internal passage. The working fluid issupplied and discharged from each plate by supply and discharge linesdedicated to each plate.

The four-stage heat exchange system also includes second stage heatexchange rack vertically aligned with the first heat exchange rack, thesecond stage heat exchange rack includes one or more open-flow heatexchange plates substantially similar to those of the first stage andvertically aligned with the plates of the first stage.

A third stage heat exchange rack, substantially similar to the first andsecond stage racks is also included and is vertically aligned with thesecond stage heat exchange rack. A fourth stage heat exchange racksubstantially similar to the first, second and third stage racks isincluded and vertically aligned with the third stage heat exchange rack.

In operation, the non-working fluid flows through the first stage heatexchange rack and surrounds each open-flow plate therein for thermalinteraction with the first working fluid flowing within the internalflow passages of each plate. The non-working fluid then flows throughthe second stage heat exchange rack for thermal interaction with thesecond working fluid. The non-working fluid then flows through thesecond stage heat exchange rack for thermal interaction with the secondworking fluid before flowing through the third stage heat exchange rackfor thermal interaction with the third working fluid. The non-workingfluid flows through the third stage heat exchange rack for thermalinteraction with the third working fluid before flowing through thefourth stage heat exchange rack for thermal interaction with the fourthworking fluid. The non-working fluid is then discharged from the heatexchange system.

Free-Flow Heat Exchange Plates:

The low temperature differential of OTEC operations (typically between35 degrees F. and 85 degrees F.) requires a heat exchange plate designfree of obstructions in the flow of the non-working fluid and theworking fluid. Moreover the plate must provide enough surface area tosupport the low temperature lift energy conversion of the working fluid.

Conventional power generation systems typically use combustion processwith a large temperature lift system such as a steam power cycle. Asenvironmental issues and unbalanced fossil fuel supply issues becomemore prevalent, Low Temperature Lift Energy Conversion (LTLEC) systems,such as the implementations of OTEC systems described herein, and whichuse renewable energy sources such as solar thermal and ocean thermal,will become more important. While conventional steam power cycles usesexhaust gas from combustion process and are usually at very hightemperatures, the LTLEC cycles use low temperature energy sourcesranging from 30 to 100 degrees C. Therefore, the temperature differencebetween the heat source and heat sink of the LTLEC cycle is much smallerthan that of the steam power cycle.

FIG. 14 shows the process of a conventional high temperature steam powercycle in a pressure-enthalpy (P-h) diagram. Thermal efficiency of thesteam power cycle is in the range of 30 to 35%.

In contrast, FIG. 15 shows the P-h diagram of an LTLEC cycle, such asthose used in OTEC operations. Typical thermal efficiency for an LTLECcycle is 2 to 10%. This is almost one-third to one-tenth that of aconventional high temperature steam power cycle. Hence, an LTLEC cycleneeds much larger size heat exchangers than conventional power cycles.

The heat exchange plates described below provide high heat transferperformance and also low pressure drop in heat source and heat sinkfluid sides to limit the pumping power requirements which affect thesystem efficiency. These heat exchange plates, designed for OTEC andother LTLEC cycles, can include the following features:

1) A working fluid flow path having a mini-channel design. This can beprovided in a roll-bonded aluminum heat exchange plate and provides alarge active heat transfer area between the working and non-workingfluids;

2) A gap provided between plates and/or offsetting the roll-bond platesbetween even number and odd number plates so as to significantly reducethe pressure drop in heat source and heat sink non-working fluids. Inthis way, a relatively wide fluid flow area for heat source and heatsink fluid sides can be provided, while maintaining a relatively narrowfluid flow area for the working fluid of the power cycle;

3) A configuration of progressively changing channel numbers per passwithin the flow passages of the working fluid can reduce the pressuredrop of the phase-changing working fluid along the flow. The number ofchannels in the plate can be designed according to the working fluid,operating conditions, and heat exchanger geometry.

4) A wavy working fluid flow passages or channel configuration canenhance the heat transfer performance.

5) Within the working fluid flow channels and among parallel channels,both ends of channel's inner walls of the flow channel can be curved tosmoothly direct the fluid to subsequent channels when the flow directionis reversed, and non-uniform distances from the ends of channel's innerwalls to the side wall can be used among parallel channels.

The above features can reduce the pumping power needed in the system,and enhance the heat transfer performance.

Referring again to FIG. 11, mini-channel roll-bonded heat exchangeplates 1051 and 1052 are shown in perspective view. A cross-counter flowbetween the working fluid and the non-working fluid is provided. Whenused as an evaporator, the non-working fluid 1057 (e.g., seawater)enters the top of the plates and leaves from the bottom of the plates.The working fluid 1058 (e.g., ammonia) enters the bottom side of theplates in liquid state, and evaporates and finally becomes vapor phaseby absorbing thermal energy from the higher temperature non-workingfluid. The generated vapor 1059 leaves the plates from the top side.

FIG. 13 shows fluid flows in a side view. The working fluid flowchannels 1055 have relatively wide width w and relatively low height hin order to increase the active heat transfer area between the twofluids while reducing the volume of the entire heat exchange plate. Thewidth w of the channels can range between about 10 and about 15 mm(e.g., more than 11 mm, more than 12 mm, more than 13 mm, less than 14mm, less than 13 mm, and/or less than 12 mm). The height h of thechannels can range between about 1 and about 3 mm (e.g., more than 1.25mm, more than 1.5 mm, more than 1.75 mm, more than 2 mm, less than 2.75mm, less than 2.5 mm, less than 2.25 mm and/or less than 2 mm). Thespacing between channels can be between about 4 and about 8 mm (e.g.,more than 4.5 mm, more than 5 mm, more than 5.5 mm, less than 7.5 mm,less than 7 mm, and/or less than 6.5 mm). The roll-bonded plates arearranged in an even plate 1051 and odd plate 1052 distribution withoffset working fluid flow passages 1055 in order to provide a smoothflow path for the non-working fluid 1057 and provide a wider non-workingfluid flow area than the working fluid flow area in working fluid flowchannels 1055. This arrangement reduces the pressure drop in the heatsource and heat sink fluid sides.

FIG. 16 illustrates an undulating or wavy working fluid flow pathdesigned to enhance the heat transfer performance of the plate.

FIG. 17 illustrates an embodiment of a heat exchange plate with twoinlets receiving working fluid 1058 and two outlets discharging heatedor cooled fluid 1059. The internal flow paths within each open-flowplate are arranged in an alternating serpentine pattern so that the flowof the working fluid is substantially perpendicular or cross-flow to theflow direction of the non-working fluid. In addition, the progression ofthe working fluid through the serpentine patter can be generallyparallel to the flow of the non-working fluid or opposite the directionof flow of the non-working fluid. In some embodiments, flow distributionbetween channels can be improved by the use of guide vanes. FIG. 18illustrates an embodiment of a heat exchange plate in which an area 1710of varying space in the flow path 1701 is provided to even the flowdistribution among parallel channels 1705. Furthermore, both ends 1715of the channel's inner walls 1712 are curved to smoothly direct thefluid to subsequent channels when the flow direction is reversed, andnon-uniform distances from the ends of channel's inner walls 1712 to theside wall 1702 can be used among parallel channels. These guide vanesand varying flow path dimensions can be implemented in heat exchangeplates such as, for example, the heat exchange plates shown in FIGS. 17,19A and B, and 20A and B.

In some embodiments, it has been found that the working fluid changesits phase from liquid to vapor along the flow path, and consequently theworking fluid pressure drop will increase significantly if the same flowpassage area is used throughout the entire heat exchange plate like. Inorder to reduce the fluid-pressure drop increase along the flowassociated with its vapor quality change, the number of parallel flowpassages per pass can be increased along the flow path of the workingfluid.

FIGS. 19A and 19B illustrate a pair of heat exchange plates 1905, 1910implementing this approach in an evaporator. The heat exchange plate1905 in FIG. 19A has two inlets 1911 which each feed into twomini-channels 1912. The mini-channels 1912 extend along the plate in aserpentine fashion that is similar to the channels of the heat exchangeplate shown in FIG. 17. However, in the heat exchange plate shown inFIG. 19A, the flow from two mini-channels feeds into three mini-channelsat a first transition point 1914. The flow from the three mini-channelsfeeds into four mini-channels at a second transition point 1916. As theheat exchange plate includes two separate, complementary flow paths,these expansions result in eight mini-channels which discharge throughfour outlets 1918.

The four outlets 1918 of the heat exchange plate 1905 are hydraulicallyconnected to the four inlets 1920 of heat exchange plate 1910 shown inFIG. 19B. The flow from four mini-channels feeds into five mini-channelsat a third transition point 1922. The flow from the five mini-channelsfeeds into six mini-channels at a fourth transition point 1924. As thisheat exchange plate also includes two separate, complementary flowpaths, these expansions result in twelve mini-channels which dischargethrough six outlets 1926. Connecting the heat exchange plates 1905, 1910in series provides the equivalent of a single long heat exchange platebut is easier to manufacture.

The plates 1905, 1910 have a length L of between about 1200 mm and 1800mm (e.g., more than 1300 mm, more than 1400 mm, more than 1450 mm, morethan 1475 mm, less than 1700 mm, less than 1600 mm, less than about 1550mm and/or less than 1525 mm). The width W of the plates can rangebetween about 250 and about 450 mm (e.g., more than 275 mm, more than300 mm, more than 325 mm, more than 350 mm, less than 425 mm, less than400 mm, less than 375 mm and/or less than 350 mm).

In some embodiments, different size plates and different numbers ofinlets and outlets are used to provide the desired heat exchange areaand expansion/contraction characteristics. For example, the pairedplates 1905, 1910 are sized in part based on the limitations of thecurrent vendor. In some embodiments, a single plate will replace thepaired plates 1905, 1010 thus removing the need for the outlets 1920 andinlets 1918 that are used to transfer working fluid from plate 1905 toplate 1910. The larger plates can have a length L of between about 2700mm and 3300 mm (e.g., more than 2800 mm, more than 2900 mm, more than2950 mm, more than 2975 mm, less than 3200 mm, less than 3100 mm, lessthan about 3050 mm and/or less than 3025 mm). The larger plates can havea width W between about 550 and about 850 mm (e.g., more than 575 mm,more than 600 mm, more than 625 mm, more than 650 mm, less than 825 mm,less than 800 mm, less than 775 mm and/or less than 750 mm). In someembodiments, a single larger inlet 1918 replaces the 2 inlets of plate1905 and feeds working fluid to all four mini-channels 1912. Because theinlets 1918 and outlets 1920 can be sources of head losses that decreasethe efficiency of the heat exchange plates, reducing the number ofinlets 1918 and outlets 1920 will reduce the overall pumping requirementand, thus parasitic load, of a given OTEC system. The flow through heatexchange plates 1905, 1910 is described for an evaporator. The heatexchange plates 1905, 1910 could also be used in a condenser. However,the flow of fluid through a condenser would be the reverse of the flowdescribed for the evaporator.

Some heat exchange plates include meandering mini-channels which canincrease residence time for the working fluid (e.g., ammonia) passingthrough the heat exchange plates as well as providing additional surfacearea for heat transfer. FIGS. 20A and 20B illustrate a pair of heatexchange plates 2005, 2010 that is generally similar to the heatexchange plates 1905, 1910 shown in FIGS. 19A and 19B. However, themini-channels of heat exchange plates 2005, 2010 include a meanderingpattern. Based laboratory testing and numerical modeling, the heatexchange plates 2005, 2010 including a sinusoidal meandering pattern areestimated to provide the same heat exchange as plates 1905, 1910 with anapproximately 10% reduction in the number of plates.

Both plates 1905, 1910 and plates 2005, 2010 include channels arrangedin relatively sinusoidal curve patterns. These patterns appear toprovide several advantages. The relatively sinusoidal curve patternscause the water flow over the plates to take a more turbulent and longerpath between the plates enabling the working fluid (e.g., ammonia) sideto theoretically extract more thermal energy from the water. Moreover,the sinusoidal flow patterns are configured such that the plates can beturned in opposite directions or staggered (e.g., alternating left andright) so that the inlet and outlet fittings do not interfere with eachother.

Heat exchange plates incorporating the various features discussed abovecan be manufactured using a roll-bonded process. Roll bonding is amanufacturing process by which two metal plates are fused together byheat and pressure then expanded with high pressure air so that flowchannels are created between the two panels. A carbon-based material isprinted on the bottom panel in the desired flow pattern. A second panelis then laid atop the first panel and the two panels are then rolledthrough a hot rolling press where the two panels are fused everywhereexcept where the carbon material is present. At least one channel isprinted to the edge where a vibrating mandrel is inserted between thetwo plates creating a port into which pressurized air is injected. Thepressurized air causes the metal to deform and expand so that channelsare created where the two plates are prevented from fusing together.There are two ways that roll bonding can be done: continuous, whereinthe metal is run continuously through hot roll presses off rolls ofsheet metal; or discontinuous wherein precut panels are individuallyprocessed.

In a prototype, two metal sheets, each approximately 1.05-1.2 mm thick,1545 mm long, and 350 mm wide, were roll-bonded together to form plates.Channels, in the patterns shown in FIGS. 19A and 19B, were formedbetween the joined metal sheets by blow-molding. The channels wereformed with a width w of between 12-13.5 mm and a height h of about 2mm. The plates exhibit good heat exchange properties using ammonia asthe working fluid and water as the non-working fluid.

Heat Exchange Plate Examples

Heat exchangers are large and costly components for OTEC powerproduction and OTEC water desalination systems. Their size andperformance dominate all other aspects of the plant design. Althoughtheoretical heat transfer coefficients range from 500 to 3,500 Btu/hft2R, the actual heat transfer coefficients of the state-of-the-art plateheat exchangers (Pheat exchanger) under given operating condition for anOTEC system are only 215-383 Btu/ft2 hrR. This poor performance wasmainly due to unbalanced flow area ratio between ammonia and watersides.

Laboratory and modeling studies were performed and it was found thatwhen plates similar to the embodiment depicted in FIG. 19A were arrangedas an evaporator having a warm water inlet temperature of 24.3 degreesC. and an ammonia inlet temperature of 10.7 degrees C., with distancebetween plates of approximately 6.3 to 6.4 mm, the ammonia quality wasmaintained between 0.0 and 0.5 and the following performance resultswere observed as listed in Table 1 below:

TABLE 1 HX Plate Performance Results under Varying Flow ConditionsProperty Test 1 Test 2 Test 3 Ai (m²) 3.232 3.232 3.232 Ui (W/m²k) 38423822 4053 Ao (m²) 5.471 5.471 5.471 Uo (W/m²k) 2270 2258 2395 DP water(KPa) 3.85 3.94 4.06 DP per length Water (KPa/m) 2.5 2.6 2.6 DP Ammonia(KPa) 10.4 13.2 10.9 MFR Water (kg/s) 11 12.0 13.0 MFR Ammonia (g/s)31.0 31.0 31.0 Heat Tranfer Coefficient Range of Range of Range of(BTU/ft²hrR) 360 to 440 360 to 435 370 to 460 Pressure Drop (psi/ft)Range of Range of Range of 0.109 to 0.111 to 0.115 to 0.113 0.116 0.119.In a comparison of rollbonded heat exchange plates and traditionalgasket heat exchange plates, the water flow rate for the rollbondeddesign was almost quadrupled while the water side pressure drop was onlyone quarter that of the gasketed heat exchange plate design. The Uivalue was also enhanced by approximately 72%. Higher Ui values can beobtained when the water flow rate is increased.

In further examples, a combined laboratory/numerical modeling study wasperformed to assess novel heat exchanger plate design concepts in orderto reduce heat exchanger surface area and the associated costs for aplanned 100 MW OTEC plant. Initial assessments of novel heat exchangerwere performed and were followed by testing of several prototype heatexchangers.

Planned Plant Design and Operating Conditions

The study estimated required heat exchanger plate surface area for aplanned 100 MWe OTEC based generally on the approach described withrespect to FIGS. 1-4. The layout of evaporators in the OTEC system forthe planned 100 MWe plant consists of 16 chambers as shown in FIG. 21A.Each chamber contains 4 stacks with each stack including 3 modules withdimensions 3 ft (width) 28 ft (depth) by 10 ft (height) as shown inFIGS. 21B-21D. As used in the study, a “cartridge” indicates a singleplate; a “cassette” indicates 2 cartridges forming ammonia flow channel;a “module” indicates 28′ assemblies of cassettes; a “stack” indicates afour-stage module; and a “chamber” includes 4 stacks.

The designs were assessed in terms of the plate surface area and volumewith a high heat transfer coefficient and low pressure drop desired.Each of the four-stage cycles producing 25 MWe are identical in designand is found to require approximately 10,080 ft³ of the evaporatorvolume is required.

Assumed operating conditions of the second-stage of a four-stage cyclefor 25 MWe were as follows: warm water inlet temperature—76.3° F.; warmwater outlet temperature—73.4° F.; evaporating pressure—927.3 kPa; LogMean Temperature Difference (LMTD)—1.7 K; and warm water flowrate—1,100,000 gpm.

Initial Designs

Five different designs were suggested and investigated. The flowpatterns and pass types are described in Table 2 below. While asingle-pass design in ammonia-side was applied to two base cases, amulti-pass design in ammonia-side was applied to three designs (DesignA, Design B1, and Design B2). The difference between the Design A andthe Design B is the ammonia channel direction. While ammonia channelsare vertically oriented in the Design A, they are horizontally orientedin the Design B so that the flow direction is in cross flow with respectto the water flow.

Cartridge size, numbers, and heat exchanger volumes were calculatedwhile maintaining the pressure drop per each stage at 2 psi.

TABLE 2 Flow pattern and pass type of heat exchangers Base case - Basecase - Design- Design- Item low area ratio high area ratio A B1Design-B2 Pass Single Single Multi Multi Multi Flow Counter CounterMixed Cross Cross pattern

FIG. 22A shows a schematic of the base case with low flow area ratio.The flow area ratio (the ratio between water flow area and ammonia flowarea) is set to 1.47. Ammonia (shown as the dark lines in the figure)flows upward through the ammonia channels and water flows downwardthrough the water channels (the gaps between ammonia channels). FIG. 22Bshows a cross-sectional view of the ammonia channel.

FIG. 23A shows a schematic of the base case with high flow area ratio.The flow area ratio is set to 7.14. While ammonia flows upward throughthe ammonia channels, water flows downward through the water channels sothat two fluids form the counter flow heat transfer configuration. FIG.23B shows a cross-sectional view of the ammonia channel. In this design,the water channel width was increased in order to increase the flow arearatio.

FIG. 24A shows a schematic of Design A. While ammonia flows through theammonia channels so that it flows up and down, water flows only downwardover the ammonia channels. FIG. 24B shows a cross-sectional view of theammonia channel. In this design, the ammonia mass flux is increased ascompared to the base cases. In this design, two fluids form counter flowand parallel heat transfer configurations alternatively.

FIG. 25A shows a schematic of Design B1. In this design, ammonia flowchannel direction is rotated in 90 from the Design A so that two fluidsform a cross-counter flow heat transfer configuration. In this design,ammonia flows through the ammonia channels horizontally and graduallymoves up, while water flows downward over the ammonia channel so thatthe potential issue of vapor stagnation on the top horizontal portion ofthe channels in the Design A is addressed. FIG. 25B shows across-sectional view of the ammonia channel.

FIG. 26A shows a schematic of the Design B2 which is also cross-counterflow as used in the Design B1. FIG. 26B shows a cross-sectional view ofammonia channel. The main difference between the Design B1 and Design B2is the ammonia flow channel area. The Design B2 has a reduced ammoniaflow channel area as compared to the Design B1. In order to control theammonia mass flux in the Design B2, the number of ammonia passes percassette is doubled by having two inlets and two outlets (see FIG. 26A).

Table 3 shows a system level comparison of the five different initialdesigns of roll-bonded heat exchangers. For the base designs, thecartridge size was extremely large due to the relatively small overallheat transfer coefficient (U) mainly due to the low ammonia heattransfer coefficient (HTC). Therefore, the base designs needed a largeplate area to provide the required total heat transfer capacity. Whilethe heat exchanger surface areas of the base cases are in anunacceptable range, the heat exchanger surface areas of Design A, DesignB1 and Design B2 are less than one million ft², which is in theacceptable range for the planned OTEC system. In Design A heatexchangers, the ammonia vapor can possibly be trapped in the horizontalportion of channels potentially causing significant problems. The DesignB2 heat exchangers provided the required total heat transfer capacitywith a smaller footprint and lower cost than the Design B1 heatexchangers. Accordingly, the Design B2 heat exchanger design was chosenas a final design.

TABLE 3 Comparison of Initial Designs Base case - Base case- high lowflow flow area Design- Design- Parameters Unit area ratio ratio Design-AB1 B2 Water-side Btu/ft²hrR 545 480 941 1,171 1,191 HTC AmmoniaBtu/ft²hrR 116 379 1,870 3,917 7,619 HTC U Btu/ft²hrR 95 210 611 871 990Cartridge ft × ft 3 × 34.58 3 × 67.05 3 × 10.67 3 × 10 3 × 9.5 Size No.of EA 40,778 28,692 20,584 16,300 16,296 cartridge heat Ft³ 83,272113,616 12,970 10,585 7,391 exchanger Volume Plate Ft² 4,767,4656,198,791 742,576 483,959 464,550 surface area Comments Unacceptableresults Infeasible OK OK design * Water-side HTC was obtained from CFD.

Prototype Evaporator Design for Laboratory Testing

The evaluation of the initial design alternatives led to the selectionof Design B2 as the flow pattern to be used in prototype heat exchangersfor laboratory testing. Several additional designs based on the DesignB2 were also tested.

The original cartridge size was 3 by 10 ft. However, in order toevaluate these heat exchangers in the laboratory test facility, the heatexchanger scale has to be cut down due to the height limitations of thetest facility.

The typical way to downscale an object undergoing hydraulic testing isusing “similarity analysis”. This method relies on maintainingdimensional aspect ratios as well as maintaining thermal properties anddimensionless numbers such as Reynolds number and Prandtl number.However, the similarity analysis approach could not be used for thisapplication as the plate thickness for the heat exchangers is fixed at1.5 mm based on a minimum required 1.0 mm for manufacturing and ammoniagap size cannot be reduced further. Accordingly, the followingalternative method was used.

As mentioned above, the original cartridge size was 3 by 10 ft. The testcartridge size was selected as a quarter size of the original cartridgesize based on test facility constraints (see FIGS. 27A and 27B). For thetest heat exchanger, the water temperature change across the heatexchanger was reduced proportionally to the cartridge height reduction,and the ammonia vapor quality change was also reduced proportionally atthe same evaporating temperature. Ammonia and water Reynolds numberswere kept constant.

The prototype heat exchange was designed with four cassettes to minimizeedge effects.

FIG. 28 illustrates pattern 1 which included 40 ammonia channels and 2dummy channels. The figure is rotated counter-clockwise by 90 degrees,so that left hand side of the figure is the top of the heat exchangerand right hand side is the bottom. Ammonia entered though the bottom andexited from the top. The ammonia mass flux was designed to be constantalong the flow path. Two dummy channels were designed to form waterchannels. Pressure in the ammonia channel was measured through the dummychannels. The dummy channels and associated pressure transducers are fortesting purposes and will not be included in production models.

FIG. 29A illustrates pattern 2 which included four ammonia channelsconnected together. For the better flow distribution among parallelammonia channels, the cross-sectional area of header section is designedto be varied along the flow as previously discussed with respect to FIG.18. The ammonia mass flux was designed to be uniform throughout the heatexchanger. In an alternate ammonia channel design, the corners arerounded, and guide vanes are applied in order to minimize the pressuredrop due to sharp corners as shown in FIG. 29B.

FIG. 30A illustrates pattern 3 which included an increasing number ofammonia channels along the flow path such that, as the ammonia vaporquality increases, the ammonia pressure drop increases as well.Therefore, the ammonia mass flux was designed to be varied to the vaporquality with high ammonia mass flux at low vapor quality and low ammoniamass flux at high vapor quality. In an alternate ammonia channel design,the corners are rounded, and guide vanes are applied in order tominimize the pressure drop due to sharp corners as shown in FIG. 29B.

FIG. 31 illustrates pattern 4. Several surface enhancements such asdimples, offsets, and a wavy pattern were considered to improve theperformance of the heat exchanger. These kinds of surface enhancementoptions are anticipated to increase the system heat transfercoefficients but also to increase the pressure drop. When the pressuredrop was regulated to the target value, the heat transfer coefficientvalue can be higher or lower than that of pattern 3. A wavy pattern wasapplied in the pattern 4 and its performance characteristics wereevaluated through the experimental test.

FIG. 32 illustrates pattern 5 which included an offset pattern appliedto the ammonia channel as an option for the surface enhancement. As sameas for the pattern 4, the relation between the heat transfer coefficientand pressure drop were investigated through the experimental test.

The performance and physical properties of suggested patterns areinvestigated for the OTEC system. Table 3 shows the summary of differentpattern designs for one evaporator of four-stage cycle for 25 MWe. Thewater channel gap and number of plates were calculated such that the HXstack depth does not exceed 28′ limitation.

For the pattern 1, the U value of the evaporator was calculated at 990Btu/ft2 hrR due to the high ammonia heat transfer coefficient. However,the ammonia-side pressure drop was predicted as 975 psi, which is evenlarger than the absolute working pressure. Thus, the ammonia pressuredrop would be the limiting factor to make this pattern feasible.

The pattern 2 was suggested with the intention of reducing the largeammonia-side pressure drop of the pattern 1. Ammonia-side pressure dropthrough the evaporator was reduced to less than 50 kPa by regulating theammonia mass flux. The U value was calculated as 660 Btu/ft2 hrR. Theplate number and evaporator volume of one of four-stage cycle for 25 MWwere calculated as 22,250 EA and 9,526 ft3, respectively.

The U value of the pattern 3 was calculated as 629 Btu/ft2 hrR, which islower than that of the pattern 2 by 5%, while the ammonia-side pressuredrop was calculated as 38 kPa that is smaller than that of the pattern 2by 16%.

For patterns 4 and 5, water and ammonia heat transfer coefficient wereassumed to be increased by 5% as compared to the pattern 3. When thesevalues are increased, the U value increases too. This results in asmaller active heat transfer area needed, so that plate numbers can bereduced and, thus, the water and ammonia mass fluxes can be increased.While the water-side HTC is increased by 5%, the ammonia-side HTC isincreased by 13.5% due to the increased ammonia mass flux and surfaceenhancement. Finally, the U value was estimated to 688 Btu/ft2 hrR.

Table 5 shows the number of plates and evaporator volume for the pattern2 design for the 100 MW OTEC system as an example.

TABLE 4 Calculated Heat Exchanger Properties Parameter Unit Pattern 1Pattern 2 Pattern 3 Pattern 4/5 Water HTC W/m²K 6,760 6,339, 6,380 6,699Btu/ft²hrR 1,191 1,116 1,124 1,180 Ammonia HTC W/m²K 43,261 9,768 8,83310,026 Btu/ft²hrR 7,619 1,720 1,556 1,766 U value W/m²K 5,622 3,7463,572 3,909 Btu/ft²hrR 990 660 629 688 Ammonia mass flux kg/m²s 588.786.3 134 -> 58 149 -> 64 Lb/ft²s 121 18  27 -> 12  31 -> 13 Ammoniapressure kPa 6,032 45.5 38.2 46 drop in evaporator Psi 875 6.6 5.5 6.7Plate numbers EA 16,296 22,248 23,784 21,520 Evaporator volume ft³ 7,3919,528 9,956 9,320 (one of four-stage cycle for 25 MWe) Plate size ft byft 3 by 9.5 3 by 10 3 by 10 3 by 10 Cassette numbers per EA/ft 24.3 33.135.4 32.0 unit length Comments — Infeasible Feasible OK Estimated Designdesign *: Water-side HTC was obtained from CFD.

TABLE 5 Estimated Evaporators with Pattern 2 Designs for 100 MWe OTECPlant 16 Chambers One One One 4 chambers for Parameters Module StackChamber for 25 MWe 100 MWe No. of plate 1,854 5,562 22,248 88,992355,968 (EA) No. of 927 2,781 11,124 44,496 177,984 cassette (EA)Evaporator 794 2,382 9,528 38,112 152,448 volume (ft³)

Table 6 shows the measured heat exchanger properties. Ammonia-side HTCwas calculated as an average value between Kandlikar (1990) and Shah(1982) correlations, and ammonia-side pressure drop was calculated as anaverage value between Lockhard and Martinelli (1949) and Friedel (1980)correlations. Eight cartridges with dimensions of 1.5 by 5 ft were to beused for the prototype evaporators.

TABLE 6 Measured Heat Exchanger Properties Parameter Unit Pattern 1Pattern 2 Pattern 3 Pattern 4/5 Water HTC W/m²K 6,760 6,339, 6,380 6,699Btu/ft²hrR 1,191 1,116 1,116 1,180 Ammonia HTC W/m²K 34,297 9,301 8,3378,985 Btu/ft²hrR 6,040 1,638 1,468 1,582 U value W/m²K 5,437 3,675 3,5153,740 Btu/ft²hrR 958 647 619 659 Ammonia mass flux kg/m²s 440.8 81 162-> 54 167 -> 56 Lb/ft²s 90 16.6  33 -> 11  34 -> 11 Ammonia pressure kPa1,908 25.1 20 21.7 drop in evaporator Psi 277 3.6 3 3.15 Plate size ftby ft 1.5 by 5 1.5 by 5 1.5 by 5 1.5 by 5 Plate numbers EA 8 8 8 8Comments — One long 4 channels Connected Surface channel connectedchannel Enhanced numbers increased *: Water-side HTC was obtained fromCFD.

Computational Fluid Dynamics Simulation on Water-Side

CFD simulation was conducted on the water-side of rollbond heatexchanger, in order to investigate the heat transfer coefficient andpressure drop characteristics. Several different designs were evaluated.

FIG. 33 shows the dual gap design in ammonia side. For the base casewith high flow are ratio, low fin efficiency resulted in needing manynumber of plates and huge heat exchanger volume. Therefore, in order toincrease the fin-efficiency high, the concept of increasing ammonia gapcame up and evaluated. The detailed geometry of the proposed concept isshown in Table 7.

TABLE 7 Detailed geometry of dual gap design in ammonia side Parameter Ab c D t Unit Mm mm mm mm mm Value 10 0.5 0.75 0.5 0.5

Table 8 shows the CFD result of dual gas design in ammonia side.Basically, it shows high water HTC and fin efficiency. However, thestress analysis results showed that the design could not stand theammonia pressure since the gap (d) was relatively small and the width(a) was large. Therefore, it was concluded that this design wasinfeasible.

TABLE 8 CFD result of dual gap design in ammonia side Water HTC DP Massflux Efficiency Re SI unit W/m2K Pa/m kg/m²s — — 10,721 5,682 2,915 94%76,242 IP unit BTU/hft²° F. Psi/ft Lb/ft²s — — 1,888 0.25 597 94% 76,242

The detailed geometry of the Design B1 is shown in FIG. 34. This is across counter flow design. Several different gap sizes were evaluatedthe heat exchanger. Since it is a cross counter flow design, the channelpattern was designed and then the ammonia-side heat transfer coefficientand pressure drop were determined. Then, the heat exchangers withdifferent water gap sizes were evaluated for the water-side.

The CFD simulation results of different gap sizes are shown in Table 9.Since cartridge size was fixed to 3 by 10 ft, the pressure drop wasmaintained at 0.2 psi/ft. Water heat transfer coefficient ranged from1,000 to 1,200 Btu/ft²Rhr.

TABLE 9 CFD result of Design B1 Water Mass Fin HTC DP flux Efficiency Reg = SI unit W/m²K Pa/m kg/m²s — — 5.3 mm 5,974 4,575 799 96%  4,400 IPunit BTU/ psi/ft Lb/ft²s — — hft² ° F. 1,052 0.2 164 96%  4,400 g = SIunit W/m²K Pa/m kg/m²s — — 8.7 mm 6,009 4,614 860 96% 10,652 IP unitBTU/ psi/ft Lb/ft²s — — hft² ° F. 1,058 0.20 176 96% 10,652 g = SI unitW/m²K Pa/m kg/m²s — — 9.4 mm 6,757 4,542 1,108   96% 14,782 IP unit BTU/Psi/ft Lb/ft²s — — hft² ° F. 1,190 0.2 227 96% 14,782

The detailed geometry of the Design B2 is shown in FIG. 35. Compared tothe Design B1, the ammonia channel gap height of Design B2 was reducedfrom 6.5 to 20.0 mm. This approach reduces the water-side pressure drop.Therefore, the water mass flux can be increased in order to enhance thewater-side HTC.

Table 10 shows the CFD result of the Design B2 while the pressure dropwas maintained at 0.2 psi/ft for all cases. The water-side HTC rangedfrom 1,200 to 1,400 Btu/ft²Rhr.

TABLE 10 CFD result of Design B2 Water Mass Fin HTC DP flux EfficiencyRe g = 5 mm SI unit W/m²K Pa/m kg/m²s — — 6,663 4,585 1,732 94% 12,349IP unit BTU/ psi/ft Lb/ft²s — — hft² ° F. 1,173 0.203 355 94% 12,349 g =6 mm SI unit W/m²K Pa/m kg/m²s — — 7,131 4,593 1,936 94% 16,520 IP unitBTU/ psi/ft Lb/ft²s — — hft² ° F. 1,256 0.203 397 94% 16,520 g = 7 mm SIunit W/m²K Pa/m kg/m²s — — 7,488 4,518 2,096 94% 20,829 IP unit BTU/Psi/ft Lb/ft²s — — hft² ° F. 1,319 0.2 429 94% 20,829 g = 8 mm SI unitW/m²K Pa/m kg/m²s — — 7,948 4,454 2,226 94% 25,238 IP unit BTU/ Psi/ftLb/ft²s — — hft² ° F. 1,400 0.197 456 94% 25,238

Additional OTEC Features:

In an exemplary implementation of an OTEC power plant, an offshore OTECspar platform includes four separate power modules, each generatingabout 25 MWe Net at the rated design condition. Each power modulecomprises four separate power cycles or cascading thermodynamic stagesthat operate at different pressure and temperature levels and pick upheat from the sea water system in four different stages. The fourdifferent stages operate in series. The approximate pressure andtemperature levels of the four stages at the rated design conditions(Full Load—Summer Conditions) are:

Turbine inlet Condenser Pressure/Temp. Pressure/Temp. (Psia)/(° F.)(Psia)/(° F.) 1^(st) Stage 137.9/74.7  100.2/56.5 2^(nd) Stage132.5/72.4 93.7/53 3^(rd) Stage 127.3/70.2   87.6/49.5 4^(th) Stage122.4/68   81.9/46

The working fluid is boiled in multiple evaporators by picking up heatfrom warm sea water (WSW). Saturated vapor is separated in a vaporseparator and led to an ammonia turbine by STD schedule, seamless carbonsteel pipe. The liquid condensed in the condenser is pumped back to theevaporator by 2×100% electric motor driven constant speed feed pumps.The turbines of cycle-1 and 4 drive a common electric generator.Similarly the turbines of cycle-2 and 3 drive another common generator.In some embodiments, there are two generators in each plant module and atotal of 8 in the 100 MWe plant. The feed to the evaporators iscontrolled by feed control valves to maintain the level in the vaporseparator. The condenser level is controlled by cycle fluid make upcontrol valves. The feed pump minimum flow is maintained byrecirculation lines led to the condenser through control valvesregulated by the flow meter on the feed line.

In operation, the four (4) power cycles of the modules operateindependently. Any of the cycles can be shut down without hamperingoperation of the other cycles if needed, for example in case of a faultor for maintenance. Such partial shut downs will reduce the net powergeneration of the overall power module.

The system requires large volumes of seawater and includes separatesystems for handling cold and warm seawater, each with its pumpingequipment, water ducts, piping, valves, heat exchangers, etc. Seawateris more corrosive than fresh water and all materials that may come incontact with it need to be selected carefully considering this. Thematerials of construction for the major components of the seawatersystems will be:

Large bore piping: Fiberglass Reinforced Plastic (FRP)

Large seawater ducts & chambers: Epoxy-coated carbon steel

Large bore valves: Rubber lined butterfly type

Pump impellers: Suitable bronze alloy

Unless controlled by suitable means, biological growths inside theseawater systems can cause significant loss of plant performance and cancause fouling of the heat transfer surfaces leading to lower outputsfrom the plant. This internal growth can also increase resistance towater flows causing greater pumping power requirements, lower systemflows, etc. and even complete blockages of flow paths in more severecases.

The Cold Sea Water (“CSW”) system using water drawn in from the deepocean should have very little or no bio-fouling problems. Water in thosedepths does not receive much sunlight and lacks oxygen, and so there arefewer living organisms in it. Some types of anaerobic bacteria may,however, be able to grow in it under some conditions. Shock chlorinationwill be used to combat bio-fouling.

The Warm Sea Water (“WSW”) system handling warm seawater from near thesurface will have to be protected from bio-fouling. It has been foundthat fouling rates are much lower in tropical open ocean waters suitablefor OTEC operations than in coastal waters. When necessary, chemicalagents can be used to control bio-fouling in OTEC systems at very lowdoses that will be environmentally acceptable. Dosing of small amountsof chlorine has proved to be very effective in combating bio-fouling inseawater. Dosages of chlorine at the rate of about 70 ppb for one hourper day, is quite effective in preventing growth of marine organisms.This dosage rate is only 1/20th of the environmentally safe levelstipulated by EPA. Other types of treatment (thermal shock, shockchlorination, other biocides, etc.) can be used from time to timein-between the regimes of the low dosage treatment to get rid ofchlorine-resistant organisms.

Necessary chlorine for dosing the seawater streams is generated on-boardthe plant ship by electrolysis of seawater. Electro-chlorination plantsof this type are available commercially and have been used successfullyto produce hypochlorite solution to be used for dosing. Theelectro-chlorination plant can operate continuously to fill-up storagetanks and contents of these tanks are used for the periodic dosingdescribed above.

The seawater conduits are designed to avoid any dead pockets wheresediments can deposit or organisms can settle to start a colony.Sluicing arrangements are provided from the low points of the waterducts to blow out deposits that may get collected there. High points ofthe ducts and water chambers are vented to allow trapped gases toescape.

The Cold Seawater (CSW) system will consist of a common deep waterintake for the plant ship, and water pumping/distribution systems, thecondensers with their associated water piping, and discharge ducts forreturning the water back to the sea. The cold water intake pipe extendsdown to a depth of more than 2700 ft, (e.g., between 2700 ft to 4200ft), where the sea water temperature is approximately a constant 40° F.The entrance to the pipe is protected by screens to stop large organismsfrom being sucked in to it. After entering the pipe, cold water flows uptowards the sea surface and is delivered to a cold well chamber near thebottom of the vessel or spar.

The CSW supply pumps, distribution ducts, condensers, etc. are locatedon the lowest level of the plant. The pumps take suction from the crossduct and send the cold water to the distribution duct system. 4×25% CSWsupply pumps are provided for each module. Each pump is independentlycircuited with inlet valves so that they can be isolated and opened upfor inspection, maintenance, etc. when required. The pumps are driven byhigh-efficiency electric motors.

The cold seawater flows through the condensers of the cycles in seriesand then the CSW effluent is discharged back to the sea. CSW flowsthrough the condenser heat exchangers of the four plant cycles in seriesin the required order. The condenser installations is arranged to allowthem to be isolated and opened up for cleaning and maintenance whenneeded.

The WSW system comprises underwater intake grills located below the seasurface, an intake plenum for conveying the incoming water to the pumps,water pumps, biocide dosing system to control fouling of the heattransfer surfaces, water straining system to prevent blockages bysuspended materials, the evaporators with their associated water piping,and discharge ducts for returning the water back to the sea.

Intake grills are provided in the outside wall of the plant modules todraw in warm water from near the sea surface. Face velocity at theintake grills is kept to less than 0.5 ft/sec. to limit entrainment ofmarine organisms. These grills also prevent entry of large floatingdebris and their clear openings are based on the maximum size of solidsthat can pass through the pumps and heat exchangers safely. Afterpassing through these grills, water enters the intake plenum locatedbehind the grills and is routed to the suctions of the WSW supply pumps.

The WSW pumps are located in two groups on opposite sides of the pumpfloor. Half of the pumps are located on each side with separate suctionconnections from the intake plenum for each group. This arrangementlimits the maximum flow rate through any portion of the intake plenum toabout 1/16th of the total flow and so reduces the friction losses in theintake system. Each of the pumps is provided with valves on inlet sidesso that they can be isolated and opened up for inspection, maintenance,etc. when required. The pumps are driven by high-efficiency electricmotors with variable frequency drives to match pump output to load.

It is necessary to control bio-fouling of the WSW system andparticularly its heat transfer surfaces, and suitable biocides will bedosed at the suction of the pumps for this.

The warm water stream may need to be strained to remove the largersuspended particles that can block the narrow passages in the heatexchangers. Large automatic filters or ‘Debris Filters’ can be used forthis if required. Suspended materials can be retained on screens andthen removed by backwashing. The backwashing effluents carrying thesuspended solids will be routed to the discharge stream of the plant tobe returned to the ocean. The exact requirements for this will bedecided during further development of the design after collection ofmore data regarding the seawater quality.

The strained warm seawater (WSW) is distributed to the evaporator heatexchangers. WSW flows through the evaporators of the four plant cyclesin series in the required order. WSW effluent from the last cycle isdischarged at a depth of approximately 175 feet or more below the seasurface. It then sinks slowly to a depth where temperature (andtherefore density) of the seawater will match that of the effluent.

Additional Aspects:

The baseline cold water intake pipe is a staved, segmented, pultrudedfiberglass pipe. Each stave segment can be 40-50′ long. Stave segmentscan be joined by staggering staves to create an interlocking seam. Pipestaves can be extruded in panels up to 52-inches wide and at least50-feet in length and can incorporate e-glass or s-glass withpolyurethane, polyester, or vinylester resin. In some aspects, the stavesegments can be concrete. Staves can be solid construction. The stavescan be a cored or honeycombed construction. The staves will be designedto interlock with each other and at the ends of the staves will bestaggered there by eliminating the use of flanges between sections ofthe cold water pipe. In some embodiments, the staves can be 40-ft longand staggered by 5-ft and 10-ft where the pipe sections are joined. Thestaves and pipe sections can be bonded together, e.g., usingpolyurethane or polyester adhesive. 3-M and other companies makesuitable adhesives. If sandwich construction is used, polycarbonate foamor syntactic foam can be used as the core material. Spider cracking isto be avoided and the use of polyurethane helps to provide a reliabledesign.

In some embodiments, the envisioned CWP is continuous, i.e. it does nothave flanges between sections.

The CWP can be connected to the spar via a spherical bearing joint. Thecold water pipe can also be connected to the spar using a combination oflifting cables and a ram or dead-bolt system.

One of the significant advantages of using the spar buoy as the platformis that doing so results in relatively small rotations between the sparitself and the CWP even in the most severe 100-year storm conditions. Inaddition, the vertical and lateral forces between the spar and the CWPare such that the downward force between the spherical ball and its seatkeeps the bearing surfaces in contact at all times. This bearing, whichalso acts as the water seal, does not come out of contact with itsmating spherical seat. Thus, there is no need to install a mechanism tohold the CWP in place vertically. This helps to simplify the sphericalbearing design and also limits the pressure losses that would otherwisebe caused by any additional CWP pipe restraining structures or hardware.The lateral forces transferred through the spherical bearing are alsolow enough that they can be adequately accommodated without the need forvertical restraint of the CWP.

Though embodiments herein have described multi-stage heat exchanger in afloating offshore vessel or platform, it will be appreciated that otherembodiments are within the scope of the invention. For example, themulti-stage heat exchanger and integrated flow passages can beincorporated into shore based facilities including shore based OTECfacilities. Moreover, the warm water can be warm fresh water,geo-thermally heated water, or industrial discharge water (e.g.,discharged cooling water from a nuclear power plant or other industrialplant). The cold water can be cold fresh water. The OTEC system andcomponents described herein can be used for electrical energy productionor in other fields of use including: salt water desalination: waterpurification; deep water reclamation; aquaculture; the production ofbiomass or biofuels; and still other industries.

All references mentioned herein are incorporated by reference in theirentirety.

Other embodiments are within the scope of the following claims.

1. A heat exchange plate comprising: front and back exterior surfacesexposed to a non-working fluid; and an interior working fluid flowchannel between the front and back exterior surfaces, comprising a firstplurality of parallel flow paths in a first direction and a secondplurality of parallel flow paths in a second direction.
 2. The heatexchange plate of claim 1, wherein the first direction is opposite andparallel to the second direction.
 3. The heat exchange plate of claim 1wherein the first and second directions of the first and second flowpaths are perpendicular to the direction of flow of the non-workingfluid.
 4. The heat exchange plate of claim 1 wherein the plate furthercomprises a first region of relatively high working fluid mass flux whena working fluid has a low vapor quality and a second region ofrelatively low working fluid mass flux when the working fluid has a highvapor quality.
 5. The heat exchange plate of claim 1 wherein theinterior flow channel has an area of varying space in fluid contact withthe first plurality of parallel flow paths and the second plurality ofparallel flow paths.
 6. The heat exchange plate of claim 5 furthercomprising one or more structural walls within the first and secondplurality of flow paths, the one or more walls being generally parallelto the flow path and terminating in the area of varying space.
 7. Theheat exchange plate of claim 6 wherein the one or more structural wallscomprises directional vane at the terminal end and oriented in thedirection of flow.
 8. The heat exchange plate of claim 6 wherein the oneor more structural walls comprises a directional vane at the proximalend of the structural wall.
 9. The heat exchange plate of claim 1wherein a flow path of the first and second plurality of flow pathscomprises a void having a cross sectional area of between 155 mm and 60mm.
 10. The heat exchange plate of claim 1 wherein the heat exchangeplate is a composite blow molded plate.
 11. The heat exchange plate ofclaim 1 wherein the heat exchange plate is aluminum.
 12. The heatexchange plate of claim 1 wherein the working fluid pressure drop acrossthe plate is about 0.2 psi/ft.
 13. The heat exchange plate of claim 1wherein the non-working fluid heat transfer coefficient ranges from 900to 1400 Btu/ft2 Rhr.
 14. The heat exchange plate of claim 1 wherein thepattern of a first plurality of parallel flow paths in a first directionand a second plurality of parallel flow paths in a second directionrepeats across the length of the heat exchange plate.
 15. The heatexchange plate of claim 14 wherein the number of flow paths in the firstand second plurality of flow paths increases as the pattern repeatsacross the length of the heat exchange plate.
 16. The heat exchangeplate of claim 14 wherein the number of flow paths in the firstplurality of flow paths increases from four flow paths per firstdirection to six flow paths per first direction.
 17. The heat exchangeplate of claim 14 wherein the number of flow paths in the firstplurality of flow paths increases from two flow paths per firstdirection to four flow paths per first direction.
 18. The heat exchangeplate of claim 1 wherein the non-working fluid is sea water.
 19. Theheat exchange plate of claim 1 wherein the working fluid is ammonia. 20.The heat exchange plate of claim 1 wherein the plate is an OTEC heatexchange plate.